A new operating range enhancementdevice combined with a casing treatment and inlet guide vanes for centrifugal compressors I Tomita, B An, T Nanbu Mitsubishi Heavy Industries, Ltd, Japan ABSTRACT This paper describes a new device combined with a casing treatment and inlet guide vanes that achieves both high pressure ratio and wide operating range of centrifugal compressors. The fixed inlet guide vanes equipped at the downstream of the recirculation bypass slit, induce a counter-rotating inflow to the impeller. Numerical calculations including whole impeller, vaneless diffuser, scroll, casing treatment and inlet guide vanes were conducted to analyze the detailed internal flow field. The calculation results indicated that counter-rotating inflow induced by the inlet guide vanes increased pressure ratio at high flow rate. On the other hand, it could also enhance the stabilizing effects of the casing treatment at low flow rate. Pressure measurements of impeller inlet and were conducted using high response pressure transducers and found that unstable phenomena like rotating stall were inhibited by the new device. Finally, the map width enhancement effects of the new device were proved by performance tests and over 70% wider map width was obtained where pressure ratio was about 2.8. The newly developed map width enhancement device combined with a casing treatment and fixed inlet guide vanes provides a wider operating range and higher pressure ratio, without any extra complex variable device. 1 INTRODUCTION Recently, demands for turbochargers are growing year by year in the background of tightening of exhaust gas and fuel consumption regulations. Especially, operating range enhancement of centrifugal compressors is difficult because of trade-off between pressure ratio and surging characteristics. However many points of surging phenomena have not been cleared yet, previous studies showed that flow separations at shroud side of impeller and unstable phenomena like rotating stall should be controlled for surging characteristics improvement(1)(2). Two compressors with different operating range width were investigated with experimental and computational flow analysis. Based on the studies, we had found one of the operating range enhancement methods utilizing a blockage effect induced by tip leakage vortex breakdown(3)(4). Furthermore, we had achieved developments of high efficiency compressors and wide range compressors by its techniques(5). Only impeller developments without any additional components were conducted in these studies. Other techniques with small trade-off between surging characteristics and pressure ratio are needed for further operating range enhancements. _______________________________________ © The author(s) and/or their employer(s), 2014 79 In this study, we developed and validated operating range enhancement effects of a casing treatment (CT) and inlet guide vanes (IGV) combination. A recirculation type CT is one of the operating range enhancement method by eliminating separation at low flow rate. However a variable IGV could control compressor performances by inducing rotational inflow at compressor inlet, its complexity of mechanical structure is critical problem. 2 TEST COMPRESSORS The test compressor is for automotive turbocharger. The compressor has a backswept impeller with 6 full blades and 6 splitter blades, vaneless diffuser and scroll. The outlet diameter of the impeller is 51mm. Figure 1 shows performance maps of a normal compressor without any devices and with a recirculation type CT. Flow rate is normalized by choking flow rate at 178,000rpm of the normal compressor. The efficiency is also normalized by maximum efficiency of the normal compressor. The surging characteristics on the normal compressor are deteriorated at pressure ratio (PR) in 2.3 or higher. Its tending is often seemed in high efficiency or high pressure compressor especially. The recirculation type CT was equipped to enhance the operating range by stabilizing the internal flow field with recirculation flow at inducer at low flow rate condition. By installing CT, efficiency was about 3.8point lower than normal compressor, the pressure ratio at low flow rate rises and was maintained good surging characteristics even 2.3 or higher pressure ratio. There was little reduction in efficiency at low flow rate by its flow improvement effect. 3.5 3.5 Deterioration of the Improvement of operating range surging characteristics 3.0 3.0 2.5 2.5 Total Pressure Ratio πc Total Pressure Ratio πc 0.99 2.0 0.98 2.0 0.95 0.97 0.95 0.93 0.93 0.90 0.90 0.85 0.85 178000rpm 178000rpm 0.80 0.80 0.75 157000rpm 1.5 0.70 157000rpm 1.5 0.75 0.70 133000rpm 133000rpm 101000rpm 64000rpm at) 20℃ 101000rpm at) 20℃ 1.0 64000rpm 1.0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Volume Flow Ratio Q0/Qmax[-] Volume Flow Ratio Q0/Qmax[-] Normal compressor CT Figure 1: Compressor performance map (measurement) Figure 2 shows a cutaway view of the compressor with the new device which has recirculation type CT and IGV. The number of the IGV and its angle inducing counter-rotating inflow to the impeller was adjusted by in-house test survey. 80 Large number of IGV increases its operating range enhancement effect, but the efficiency is deteriorated due to friction loss increase. 7 vanes was the best number for efficiency and surging characteristics in the case of this compressor. And an inner ring was installed to eliminate its tip leakage vortex and decrease its vibration. This newly developed device provides a wider operating range and higher pressure ratio, without any extra complex variable device. In this study, flow mechanisms on the normal compressor, CT and the new device (IGV+CT) were investigated. Scroll Casing treatment Recirculation flow Inlet guide vanes Flow Rotation Impeller Flow Inner ring Vaneless diffuser Figure 2: The new map width enhancement device 3 NUMERICAL AND EXPERIMENTAL METHOD In this study, we analyzed the flow stabilization mechanism in steady CFD. And the unsteady phenomena was confirmed by a pressure fluctuation measurement. 3.1 CFD scheme CFX version 14.5 was used for numerical calculation. Order to flow improvement by recirculation flow at the casing treatment, evaluation of the recirculation flow rate is important. The recirculation flow rate is determined by the pressure balance of the impeller and the casing treatment. Since distortion of the flow is generated by the circumferential pressure distributions in the scroll within the low flow rate operation in particular, numerical calculations were performed including whole impeller, IGV, casing treatment, vaneless diffuser and scroll applying the Frozen-Rotor in boundary conditions. The SST turbulence model was adopted. And the numbers of cells were about 2,440,000 for impeller, 920,000 for diffuser, 1340,000 for scroll and 7,300,000 for other elements. 3.2 Measurements scheme Pressure fluctuation measurement was conducted to detect the unsteady phenomena on these compressors. As shown in Figure 3, measurement locations are 3mm upstream from impeller leading edge. The purpose of the measurement near impeller leading edge is to detect blade passing waves, rotating stall, and 81 rotating instability, etc. Two sensors are installed in the same meridional location to detect propagating phenomena in circumferential direction. After passing through the sensor 1, the impeller passes through the sensor 2. If phenomena are propagating, a phase difference of 30 degrees is detected by sensor 1 and sensor 2. When the rotating stall are generated, the speed and the number of disturbances are estimated from the phase lag of these two sensors. Endevco 8510C sensors with sampling frequency of 250kHz were applied. It corresponds to 14 points for 1 pitch on full blades at 178,000rpm. Impeller Rotation Sensor 2 Pressure transducer 3mm 30 deg Sensor 1 Figure 3: Positions of pressure transducers 4 NUMERICAL CALCULATION RESULTS Figure 4 shows relative mach number distributions at 90% span of the impeller at 178,000rpm. Figure 4(a) shows peak efficiency point (Q/Qchoke=0.836), figure 4(b) shows peak pressure point of the normal compressor with neither CT nor IGV (Q/Qchoke=0.716) and figure 4(c) shows near surging point of the normal compressor (Q/Qchoke=0.597). 4.1 Peak efficiency point (Q/Qchoke=0.836, figure 4(a)) Shock waves from suction surface and tip leakage flow were seen in figure 4(a). Mach number distributions show that flow separations did not occur and the impeller still had loading on these compressors because velocity differences between pressure side and suction side on the inducer. Two changes in flow field were observed. One was circumferential uniformity of tip leakage flow by CT, and the other was higher acceleration on suction surface caused by counter-rotating inflow induced by IGV. 4.2 Peak pressure point of the normal compressor (Q/Qchoke=0.716, figure 4(b)) On the normal compressor, a large stall region was observed and eliminating of the blade loadings was cleared from that the velocity difference between pressure side and suction side disappeared. On the other hand, any stall did not occur on CT and IGV+CT. These results confirmed that CT and IGV+CT could stabilize the flow structure by recirculating flow at a non-design point. 82 Full blade Larger Rotation incidence angle TE Without IGV Relative velocity LE With IGV Full blade High acceleration Shock wave Splitter blade Acceleration Tip leakage flow Tip leakage flow was uniformed Velocity differences by CT. Flow by blade loading Normal compressor CT IGV + CT (a) Peak efficiency point (Q/Qchoke=0.836) No stall region No stall region Stall region No velocity difference -> lost loading Flow Large velocity differences -> High-loading Normal compressor CT IGV + CT (b) Peak pressure point of the normal compressor (Q/Qchoke=0.716) No stall region No stall region Stall region No stall region but blade loadings are Keep high not uniform blade loadings on all blades Normal compressor CT IGV + CT (c) Surging point of the normal compressor (Q/Qchoke=0.597) Figure 4: Relative mach number distribution (90% span, 178,000rpm) 83 4.3 Surging point of the normal compressor (Q/Qchoke=0.597, figure 4(c)) At the surging point of the normal compressor, a large stall region could be eliminated by CT in the same way as Q/Qchoke=0.716. Even though blade loadings were not uniform on each blade on CT, IGV+CT could keep large blade loadings at all blades. IGV+CT had a stable flow structure same as the design point even at low flow rate. The obtained results suggest the operating range enhancement effect of the new device. 4.4 Recirculation flow ratio Figure 5 shows the ratio of their recirculation flow rate ∆Q and total flow rate Q. Since Q/Qchoke=0.597 was an unstable point inherently, convergence of the analysis was not fine. The recirculation flow ratio ∆Q/Q was little changed by existence of IGV at Q/Qchoke=0.597. The recirculation flow rate on IGV+CT was 12% and that on CT was 9% at Q/Qchoke=0.716, so the recirculation flow rate increased 1.4 times that of CT. The recirculation flow is generated by pressure difference on inlet and outlet of CT. Its pressure difference was increased by counter-rotating flow by IGV because the counter-rotating flow raised the blade loading. The results obtained by numerical calculation suggest that IGV improves the flow stabilizing effect of CT because of its counter-rotating flow. 0.3 CT IGV + CT 0.2 Recirculation Flow Ratio ⊿Q/Q[-] 0.1 0.0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 -0.1 -0.2 Volume Flow Ratio Q0/Qmax[-] Figure 5: Recirculation flow rate at casing treatment (178,000rpm, CFD) 5 PRESSURE FLUCTUATION MEASUREMENT Figure 6 shows measurement results of pressure fluctuations at inlet at 178,000rpm. As figure 4, figure 6(a) shows peak efficiency point (Q/Qchoke≒0.836), figure 6(b) shows peak pressure point of the normal compressor with neither CT nor IGV (Q/Qchoke≒0.716) and figure 6(c) shows near surging point of the normal compressor (Q/Qchoke ≒0.597). The solid line is pressure fluctuation measured by the sensor 1 and dotted line is that of the sensor 2. The range of horizontal axis corresponds to about 1.2 rotations of the impeller. In the case of stall, the blade loading disappeared and blade-passing fluctuations are not clear. 84 5.1 Peak efficiency point (Q/Qchoke≒0.836, figure 6(a)) Blade passing fluctuations were detected clearly and any low frequency fluctuations were not seen at this flow rate. It seems that any unsteady phenomena did not occur and their internal flow structures was steady on the normal compressor. Pressure fluctuations on IGV+CT were slightly higher by counter-rotating inflow. 5.2 Peak pressure point of the normal compressor (Q/Qchoke≒0.716, figure 6(b)) The pressure fluctuations clearly showed that unsteady flow disturbances considered being a stall occurred on the normal compressor. On the other hand, pressure fluctuations are stable on CT and IGV+CT. The effectiveness of CT and IGV+CT was confirmed experimentally. 5.3 Surging point of the normal compressor (Q/Qchoke≒0.597, figure 6(c)) Both Stall and surging did not occur on CT and IGV+CT even at the flow rate that surging occurred on the normal compressor. It was also observed that the blade loadings were not uniform on each blade on CT. And IGV+CT could keep large blade loadings at all blades. These measurement results supported the numerical calculations. Sensor 1 Sensor 2 1 pitch of full blades Pressure waves by 2 sensors Higher fluctuations by IGV Normal compressor CT IGV + CT (Q/Qchoke=0.834) (Q/Qchoke=0.833) (Q/Qchoke=0.839) (a) Peak efficiency point Unsteady flow disturbances No disturbance No disturbance Normal compressor CT IGV + CT (Q/Qchoke=0.691) (Q/Qchoke=0.716) (Q/Qchoke=0.713) (b) Peak pressure point of the normal compressor Keep large blade loadings Blade loadings are not uniform on all blades No data because of the surging No disturbance No disturbance Normal compressor CT IGV + CT (Q/Qchoke=0.598) (Q/Qchoke=0.594) (c) Surging point of the normal compressor Figure 6: Pressure fluctuation (178,000rpm, measurement) 85 6 PERFORMANCE IMPROVEMENT The compressor performance of IGV+CT and comparison are shown in figure 7. Surging characteristics of the normal compressor were deteriorated at pressure ratio in 2.3 or higher. However the compressor with CT had 38% smaller surging flow rate at 178,000rpm, its pressure ratio was dropped over 0.13 m3/s. This pressure deterioration might be occurred by loss generation at the recirculation flow passage. On the other hand, the compressor with IGV+CT had best surging characteristics that surging flow rate decreased 20% at PR=2.3 and 50% at PR=2.8 compared with the normal compressor. Moreover, IGV+CT achieved higher pressure ratio at whole flow rate by counter- rotating flow. When operating range was defined as (Qchoke-Qsurge)/Qchoke, the operating range of IGV+CT was enhanced 76% compared with the normal compressor and enhanced 14% compared with CT at 178,000rpm. The maximum efficiency was not deteriorated by IGV because the maximum efficiency of IGV+CT was almost same as CT. 3.5 Normal Compressor CT IGV+CT IGV+CT CT 3.0 Normal 2.5 Total Pressure Ratio πc 2.0 0.95 0.93 0.90 178000rpm 0.85 0.80 0.75 1.5 0.70 157000rpm 133000rpm at) 20℃ 64000rpm 1.0 0.0 0.2 0.4 0.6 0.8 1.0 1.2 Volume Flow Ratio Q0/Qmax[-] IGV + CT Performance comparison Figure 7: Compressor performance map (measurement) 7 CONCLUSION In this study, the effects of combination of a casing treatment and inlet guide vanes were investigated by numerical calculations and measurements. In conclusion, the results are summarized as follows. (1) Counter-rotating flow generated by the inlet guide vanes makes pressure ratio high at whole flow rate. Furthermore, flow unsteadiness could be controlled by combination with casing treatment. 86 10th international conference on Turbochargers and Turbocharging. “Development of advanced centrifugal compressor for turbocharger.. GT2009-59516.. (3) Tomita.. Tomita. I. S. (2) Iwakiri.. Yamada. 2011. 2009. Furukawa. pp. K. M. ASME Paper.. “Unsteady and Three- dimensional Flow Phenomena in a Transonic Centrifugal Compressor Impeller at Rotating Stall”. 2012. Shiraishi T.... Ibaraki. REFERENCES (1) Yamada. applying control of internal unsteady flow structure”. Furukawa. S.. M. I. M. Ibaraki. GT2012-68947. Tomita I.. ASME Paper.. K. (4) Tomita.. “Feature of Internal Flow Phenomena of Centrifugal Compressor for Turbocharger with Wide Operating Range”， Gas Turbine Congress， 2011. 2012. without any extra complex variable device. S. Furukawa. (5) Ebisu M. IMechE. M. K.. Ibaraki. H. I.. “The Effect of Tip Leakage Vortex for Operating Range Enhancement of Centrifugal Compressor”... ASME Paper. Tomita. 87 . “The Role of Tip Leakage Vortex Breakdown in Flow Fields and Aerodynamic Characteristics of Transonic Centrifugal Compressor Impellers”. 135-144.. K.8. Ibaraki.(2) The newly developed map width enhancement device combined with a casing treatment and fixed inlet guide vanes provides over 70% wider operating range and higher pressure ratio compared with the normal compressor where pressure ratio is about 2. I. S. GT2011-46253.. Yamada. Furukawa... Fukushima. A one-dimensional performance model for turbocharger turbine under pulsating inlet condition H Chen1. and is coupled to the volute flow through a sliding rotor- stator interface. This model is solved by the method of characteristics. n2 Indices in incidence loss equation (17) nb Number of rotor blades P Pressure (N/m2) Q Heat transfer through the walls of flow passage (J/kg-s) R Radius (m) Re Reynolds number s Clearance between rotor backdisc and heat shield (m) Str Strouhal number t Time (s) T Temperature (oK) U Rotor peripheral (tip) speed (m/s) _______________________________________ © The author(s) and/or their employer(s). The results show that the model captures important unsteady features such as hysteresis of turbine mass flow and efficiency. 2014 113 . UK ABSTRACT This paper describes a new 1-D model of turbocharger turbine working under pulsating inlet condition. D E Winterbone2 1 National Laboratory of Engine Turbocharging Technologies. but with mass removal or addition to simulate the flow into and from the rotor. Possible improvements and extensions to the model are also discussed. The flow inside the turbine volute is treated as one- dimensional and unsteady. NOMENCLATURE A Flow passage area (m2) a Speed of sound (m/s) b Flow passage width (m) Cf Coefficient in disk friction torque expression (22) CFD Computational Fluid Dynamics Cis Isentropic expansion (spout) speed (m/s) Cp Specific heat at constant pressure (J/kg) f Pulse frequency (Hz) h Specific enthalpy (J/kg) L Wet periphery of flow passage (m) m Index in angular momentum equation (14) Mf Disk friction torque (N-m) m Mass flow rate (kg/s) N Turbine rotating speed (rev/s) n1. The flow in each rotor passage is treated as one- dimensional and unsteady. China 2 Formerly UMIST. left-running Math-lines and path-line (characteristics) Cycle mean. It has long been recognised that the performance of these turbines can be quite different from that under steady flow conditions. if Strf-p << 1. Strf-p ~ 1. The following rough guides hold. unsteady effects dominate. quasi-steady effects dominate. INTRODUCTION Turbocharging turbines for internal combustion engines often work under pulsating flow conditions. both quasi-steady and unsteady effects are important. and p is the time scale of the unsteadiness of pulsating flow. total-to-static efficiency Kinematic viscosity (m2/s) Density (kg/m3) Time scale.V Absolute velocity (m/s) W Relative velocity (m/s) x Coordinate along flow direction in the volute (m) Greek Symbols Angle between absolute and peripheral velocities Angle between relative velocity and radial direction Right-hand term of compatibility conditions along characteristics Ratio of specific heats Right-running. turbine inlet 1 Centroid of turbine housing 2 Volute exit 3 Rotor inlet (after incidence) a Ambient f Fluid m Mean p Pulse r Rotor s Steady flow u Unsteady flow w Wall 1. Yet there is still a lack of understanding of the phenomena. wall shear stress (N/m2) Azimuth angle Mass withdraw from/addition to the volute per unit length per unit time (kg/s-m) Rotor angular speed (rad/s) Subscripts 0 Stagnation or total state. Strf-p >>1. The relative importance of unsteady effects produced by pulsating flow might be estimated by a Strouhal number which is the ratio of two timescales (1): Strf-p = f/ p (1) where f is the time for fluid particles to be transported through the turbine components. 114 . and there is no simple method that can be used with confidence to predict the performance of the turbine under such conditions. steady effects may dominate. The flow in the rotor is treated as quasi-steady. (8) described a turbine model in which the volute is represented as a volume between the turbine inlet pipe and the turbine rotor entry. This time scale r is largely independent of f. and showed fairly good agreement in mass flow between predictions and experiment. 0. The Wallace model was modified by Mizumachi and his co-workers (6). The modified model was used for full admission. and simplified the rotor flow passage to a one-dimensional duct. The quasi-steady flow method may produce poorer cycle-mean performance results than the simpler steady flow method. The results showed it captured unsteady effects better than a quasi-steady model. They also developed a new model for partial admission of sector divided. and simulated by a meanline model. a four-stroke. four cylinder engine rotating at 2000rpm and a turbine housing with averaged flow passage length of 150mm and an averaged flow velocity of 100m/s. The model was later (2) applied to a mixed flow turbine. Results obtained in this work (5) showed that it is possible to trace pressure waves through the turbine albeit in a relatively crude manner. Another important time scale for turbines is the time scale for the rotor to rotate one revolution. The steady flow method uses cycle mean inlet conditions and measured steady flow performance maps. but the efficiency was overestimated. The first unsteady flow model was perhaps presented by Wallace and Adgey (4) in 1967. 115 . two-dimensional duct. turbine behaviour becomes increasingly steady. which is effectively a zero-dimensional model. Rotation of the rotor was simulated by transferring the connection of the rotor passages with the nozzles one by one at a time interval of one-sixth of a rotor revolution. nozzled. Baines et al. while the quasi-steady flow method uses the same performance maps in a quasi-steady manner to predict instantaneous performance from the instantaneous inlet conditions.0015s.1 respectively.015s. Chen and Winterbone (7) proposed a housing model that replaces the volute by an equivalent length of nozzle. In this case both unsteady and quasi-steady effects exist in the housing. quasi-steady and unsteady flow models. Existing low dimensional methods for predicting turbine performance under unsteady flow fall into three categories: steady. and applied one-dimensional unsteady flow equations to this nozzle. a typical streamline in one sector of the scroll was branched twice and the branching points were treated as 'constant pressure' junctions. but it did not show any improvement on predicted cycle-mean performance versus a steady flow method. The rotor is treated as a single. the stator and the rotor were each divided peripherally into six sections. In the new model. these would give f. In this model. p and Strf-p values of 0. the nozzled turbine housing is simulated by a convergent nozzle and the one-dimensional unsteady flow equations can then be applied to the nozzle. can predict some of the measured features of unsteady flow. so quasi-steady effects should dominate the rotor flow. Averaged flow passage length of turbine rotor is about a quarter of that of the housing. and there is evidence (2-3) that with the reduction of r or the increase of turbine speed. and 0. They showed that this approach. who reformed the equations. So another Strouhal number may be needed: Strr-p= r/ p (2) and when Strr-p is small.Take for example. twin-entry housing. The mass flow rate and the pressure waves at the turbine entries were well predicted. as applied to the volute model in Figure 1. flow continuously leaves the volute along the circumferential direction of turbine volute and enters the rotor. The unsteady. but with additional terms to take into account this mass removal/addition.0025 V1C p (Tw T1 0. The model was originally devised between 1987-88 but has not been formally published.The difficulty in modelling the unsteady flow in the volute housing is that the volute is not a simple pipe and one-dimensional gas flow equations are difficult to apply. the only difference being that the flow is a function of time and may reverse.0025 1V1 2 (6) Energy 1h01 p1 1V1h01 V h dA (7) h02 1 1 01 1 1Q t x1 A1 A1 dx1 where Q is the heat transfer through the wall. entering the volute from the rotor. Treating it as a volume ignores the gas momentum equation and will inevitably lead to error.1 Curved slot model of volute with mass removal/addition Under steady flow conditions. are: Continuity 1 1V1 1V1 dA1 (3) t x1 A1 A1 dx1 where is the mass withdrawn from or added to the volute per unit length per unit time.5 ) d Momentum 1V1 1V1 p V dA1 L1 2 2 V2 cos( 2 1 ) 1 1 sign (V1 ) w (5) t x1 A1 A1 dx1 A1 where w is the wall shear stress. see Figure 1. In the remainder of this paper. 2. one-dimensional gas flow equations (9-10). This flow leaving/entering the volute may be considered as mass removal/addition to the flow inside the volute. and is expressed as: dR1 (4) 2V2 sin 2 b2 R2 /( R1 0. The same happens under unsteady flow condition. a new one-dimensional unsteady flow model of the turbine volute housing is proposed to overcome this difficulty. and is expressed as: w 0. and is expressed as: L1 (8) Q 0. PHYSICAL MODEL OF TURBINE 2.424V1 / C p ) 2 A1 116 . so the volute flow can still be treated as one-dimensional. instantaneous static temperature is then calculated using the method of (12) for inflow to the inlet. The velocity at the inlet is calculated through the characteristic.2 Method of characteristics and boundary conditions The method of characteristics (11) is used to give necessary equations for various boundaries: dp1 a1 1 dV1 a1 1 a1 2 3 dt 2 (9) along Mach lines and . A1 A1 dx1 L1 2 11 V2 cos( 2 1 ) V1 sign (V1 ) w . and for outflow from the inlet. 1V1 dA1 (13a) 11 . (13c) For the entry region of the turbine housing from the turbine inlet flange to the tongue ( = 0o). V2 +d 2. the density at the inlet is calculated through the path line characteristic. 12 . R2 Fig. A2. The volute end ( = 360o) is simplified to a closed end with V1 = V2 = 0. 2. 1 Curved slot model of turbine volute 2. Two Mach line characteristics are used to obtain the pressure and density at this point. (13b) A1 3 ( 1)11 (h02 h01 ) V1 2 Q 1 . The instantaneous static pressure and time mean stagnation temperature at the inlet flange are assumed to be known from experimental measurements. dx1 (10) V1 a1 dt and dp1 a1 d1 3 dt (11) 2 along path line dx1 (12) V1 dt where 1 11 12 . P1 dx1 P1+dP1 1 x 1+d1 V1 1 V1+dV1 y 1 A1 A1+dA1 x R2 R1 R1+dR1 R1 y P2. equations (3) to (13) are used with = 0. 117 . The indices n1 and n2 were selected from test data. the same equations as (14) to (20) are used except that equation (17) is replaced by the compatibility condition along the path line: L (21) dp a 2 d ( 1) Q W sign (W ) w dt A V2 V3 W2 W3 2<0 2>0 U2 U3 Fig. a modification of NASA's incidence loss model (13-14) is used for inflow: W3 W2 cos n1 2 . Energy and mass conservations and the compatibility condition along the characteristic running from inner rotor toward the rotor entry are also used: 2 P3 W2 P2 2 W3 (18) 2 1 3 2 1 2 3W3 A3 2W2 sin 2 A2 (19) dp adW 2 W dA L dR L a a sign(W ) w 2 R ( 1) Q W sign(W ) w dt A dx A R A (20) where Q and w are the heat transfer term and wall shear stress of the rotor passages respectively. where m is an index and may be estimated from experimental data (m = 1 implies angular momentum conservation). as shown in Figure 2. for 2 0 . angular momentum equation V2 cos 2 V1 cos 1 R1 / R2 m (14) is applied. for outflow from the volute to the rotor (W3 > 0). 2 Velocity triangles before and after incidence at rotor inlet 118 . (17b) where W2 and W3 are the relative velocities at rotor entry before and after incidence. and following assumptions are used: P2 = P1 (15) (16) At rotor inlet. for 2 0 . 2 2 (17a) 2 2 W3 W2 1 sin n2 2 . For reversed flow from the rotor to the volute ( W3 0 ).The flow between the volute exit or station 2 and the volute central line or station 1 is assumed quasi-steady. Turbine data used was supplied by Holset and are shown in Table 1.0295m Blade angle (trailing edge shroud) 63. The disc friction is estimated by the following formula (15) M f 0.4 Gas viscosity 1. Given the fact that only a very simple friction loss model was used for the rotor passages.0133 x105N/m Ambient density a 1. The volute between = 0o to 360o was divided into nb sections with each section extending over the same angle = 2/nb.3 Rotor model and solution method Equations similar to those in (6) are used for the rotor passages.080Re (23) The Lax-Friedrichs algorithm (16) has been used to solve governing equations for internal points. 3.0171m Number of rotor blades 12 Rotor inlet radius 0. the over-predicted efficiency is encouraging. where nb is the number of rotor passages. Main input data Housing inlet area 4.5C f a 2 R3 . with wall friction and heat transfer terms included. The maximum error in the predicted mass flow is about 5%. Table 1.73 x10-3m2 Gas constant 287.163 x10-3m2 Width of volute exit (station 2) 0.71 x10-3m2 Housing end ( =0) area 0. The values of indices m. RESULTS 3.225 kg/m3 119 .1 Steady flow results Steady flow calculations were first carried out to test the model. Connection between these sections and the rotor passages is carried out in every time step.2.1 J/kg-oK Gas specific heat ratio 1. and some important losses such as clearance loss and bearing loss were not included.5o Rotor exit area 1. and Re R3 / 1/ 4 2 C f 0. and the Courant-Isaacson-Rees algorithm (17) has been applied to solve the characteristic equations for boundary points.0485m Rotor inlet area 4. 5 (22) where Mf is the friction torque.125 x10-3m2 Rotor exit radius 0. These steady state results are shown in Figures 3-4. n1 and n2 were chosen at different rotor speeds to give agreement between the predicted and measured performances. Cf is given by /( S / R3 ) 1 / 6 .5 x10-5m2/s Turbine inlet total temperature 400 oK Wall temperature 320 oK Ambient pressure Pa 1. 8 P00 / Pa 2 7 6 5 P00 / Pa 1. Waves with different shapes and periods were used to investigate the influence of such parameters as frequency. 120 . This is consistent with the influence of the Strouhal number Strr-p as discussed earlier: for the highest pulse frequency of 100Hz modelled. Increasing the pulse amplitude reduces the mass flow rate as shown in Figure 9a. are 0. The results are given in Figures 8-10. the lower the turbine efficiency tends to be.2 Unsteady flow results Computations were carried out to investigate the influence of pulsating flow on turbine performance. 0.2sin(270t) and N / T00. amplitude and shape on the turbine performance. Figure 8b also suggests that the steady flow method might overestimate as well as underestimate both the efficiency and power depending on non-dimensional turbine speed N / T0 . 167 and 0. The figure further suggests that the steady flow method might give a better prediction of turbine efficiency at higher turbine speeds. The influence of pulse shape is shown in Figure 10. The relationship between turbine efficiency and the pulse amplitude. is damped at the middle of the volute as shown in Figure 7b. and hysteresis between efficiency and blade speed ratio as obtained by Dale and Watson (20). The results are shown in Figures 5-10.m 30 . The stronger the pulsating is. The figure also suggests that a steep wave-front might result in a lower turbine efficiency. the values of Strr-p as expressed by equation (2). 4 Steady efficiency prediction 3. Figure 8 shows a slight increase of mass flow rate and a slight decrease in efficiency with increasing pulse frequency.2 4 3 2 N / T00 1 0 10 20 30 40 50 0 10 20 30 40 50 Fig. 3 Steady mass flow prediction Fig.5. Figure 7a. where unsteady flow results are plotted against steady flow results. shows a strong dependence on turbine speed. Strong pulsating of the angle at the tongue region. The turbine operating loci show hysteresis between expansion pressure ratio and non-dimensional mass flow rate. Figures 5-6 show the mass flow and efficiency characteristics under a pulsating inlet condition of P00/Pa = 1. indicated in Figures 9b. These results agree with those obtained from experiment in (12).106 respectively for the three turbine speeds in the figures.4-0. Figure 7 shows the variation of flow angle at volute exit with time at two different circumferential locations. m 2.6 1. 6 Unsteady efficiency locus 28 58 2.8 0 .2 P00 ( t ) / Pa 1.92 0.95 0.5 5.3 0 .8 1. 8 Effect of pulse frequency on mass flow and efficiency 121 .0 1 . 180 (t ) 54 56 27 26 52 N / T00.0 4.9 0. s Fig. 7 Flow angle variation with time and location P00 / Pa 1 .7 0 .2 0.2 1 1.4 0.0 6.0 1.0 3. 1.2 Fig.5 0 .6 0.96 0.3 1.9 0 20 40 60 80 100 0 20 40 60 80 100 f (Hz) (a) m u / m s (b) t s .6 0 .94 0.9 1. 6 sin( 2 ft ) 1 mu/ ms t s .9 1 . 6 0 .85 0.m 30 22 0 2 4 6 8 10 12 14 0 2 4 6 8 10 12 14 (a) =330deg (b) =180deg Fig.8 f (Hz) 0.4 0 . s 1.5 3.m 30 50 25 46 48 24 40 42 44 23 N / T00.u /t s .75 0.u / t s .5 0.5 0 .5 4. 5 Unsteady mass flow locus Fig.1 1 .05 0.5 6.1 U / C is .98 1 0.0 5.7 1.8 1.4 1. 0 s 0.3 2.90 0.8 4.8 0. equations (15) and (16) for volute exit should be replaced.u /t s.6 2.2 N / T 00 .86 0.0 (a) m u / m s (b) t s .94 0.82 0. 1. Firstly.2 0.4 0.2 0.6 0.u 0.4 0. m *105 (b) t s Fig. However. since the model was first conceived over two decades ago.6 0.m / P00. is 122 .6 3. 10 Effect of pulse shape on mass flow and efficiency 4. whether meridionally divided or circumferentially divided.0 0. secondly. Extending the model to twin-entry turbine housing. Application of the current vaneless housing model to vaned housing is straight forward.2 3.4 Steady flow P00 / Pa 1. and takes into consideration more geometrical parameters of the housing.0 0 0. DISCUSSION The model presented here treats turbine housing more rationally than in other previous models. the closed end treatment of the volute end may be replaced by a junction model allowing nonzero velocity at the end. could be used to improve the accuracy of the performance prediction.4 0. s Fig. m N / T 00 . 9 Effect of pulse amplitude X on mass flow and efficiency t s . m 0 5 10 15 20 25 30 35 0 5 10 15 20 25 30 35 (a) m T00.8 0.8 1.8 1. perhaps by adiabatic flow assumption and mass conservation between stations 1 and 2. m / P00.2 sin( 2 50t ) 0.7 X X 0.8 m T00. better rotor loss models. improvements are now appropriate and possible. Station 2 is now the inlet to the nozzles which can be treated similarly to the rotor inlet.1 0.98 1 t s.m *105 3.0 3. thirdly.9 0.5 3.1 0. perhaps adopting first the steady flow loss modelling technique.6 0.s m u/ m 1.7 0. u / t s .2 0. but in the stationary frame.6 0 0. IMechE. D. 65-76. pp 179-186. 1985. 4. J. Proc. H. 6. Baines N. IMechE. E. Pure and Appl. Vol. and Glassman A.. Sci.. S. and Allison A. Oxford Uni. variable-area stators. Greitzer E. and Rees M. Mizumachi N. Pt. 10. J. and Hirsch C. Vol. Press. On the solution of non-linear hyperbolic differential equations by finite differences. 1974. NASA Technical Note D-8063.. 9. 1. Math. S. Math. Stow P. Chen H. F. SAE Meeting Milwaukee. Woods W. Benson R. Wallace F. SAE Paper 740739. pp. pp. 243-355. Turbocharging and Turbochargers. 11. 1982. 1979. 183-195.. C... 1980. NASA Technical Report 80-C-13.. pp. 210... Dale A. pp. The thermodynamics and gas dynamics of internal combustion engines. J. Vol. and Endoh T. Sept. Wasserbauer C. 2. and Martines-Botas R. 1975. E. et al. Vol. 5. Proc. pp.. Proc. J. 16. and Winterbone D. IMechE 4th Int. P.. Wallace F. I. edited by Ucer A. 7. and between the volute and the rotor. 1977. 13-22. and Winterbone D. Proc. Dec.. Courant R. REFERENCES 1. Glassman A.. Chen H. 2011. Vol.. Paper C484/006. J. NATO ASI Series.. conf.. 184. Proc. Turbine design and application. 2 of Thermodynamics and fluid mechanics of turbomachinery. A. 1967-8. 1972-75. 10. Pt. No. Uni. 13. Off-design loss model for radial turbine with pivoting. IMechE 5th Int.. A study on performance of radial turbine under unsteady flow conditions. M. J.. The pulse flow performance and modelling of radial inflow turbines. 5. F. Martinus Nijhoff Publishers. Inc. H. Theoretical assessment of the non-steady flow performance of inward radial flow turbines. Report of Inst. Hakeem I. Vaneless diffuser turbocharger performance. pp. and Blair G. A. Paper C110/86. Private communication with Honda simulation engineers in Japan. Vol. IMechE. 2. 1. IMechE conf. Adgey J.. Modelling of a turbocharger turbine under pulsating inlet conditions.. Commun. Paper C405/008. Performance of inward radial flow turbines under non-steady flow conditions. A method to predict performance of vaneless radial turbines under steady and unsteady flow conditions. 209-18. 3H. conf. 8. Non-steady flow in a turbocharger nozzless radial gas turbine. 3.. 17. IMechE. L.. pp. An introduction to unsteady flows. 1976-6. John Wiley & Sons. edited by Horlock J. 18. 123 . 28.. Weak solution of non-linear hyperbolic equations and their numerical computation. Isaacson E. Fortran program for predicting off- design performance of radial-inflow turbines. and Glassman A. 1996. 1994. for the meridionally divided housing. 182. pp. Unsteady compressible flow with gradual mass addition and area change. Pub. For the circumferentially divided housing. and Adgey J. and Watson N. and Hoffman J. A three-way junction may be suitable for this task. D. Proc. 1954. Zucrow M. F. Pt. 14.. 7. Hajilouy-Benisi A. 1969-70. Lax P. 159-193. Gas dynamics. S. Tokyo. 397-408. 15. 12. and Yeo J. 5.. IMechE Sixth Thermodynamics and Fluid Mechanics Convention. Yoshiki H. IMechE. IMechE 3rd Int. conf. Vol. Turbocharging and Turbochargers. No. Turbocharging and Turbochargers. NASA SP-290. additional modelling is required to allow for the interaction between the two branches of the volute. Benson R.. IMechE. 1952. Vol. 22-36. the extension is simpler and more straightforward. Pure and Appl. Indus.possible. 1990. Meitner P. Commun. A. Vol. 1986-4.. The shaft / bush interface is a critical tribosystem within the turbocharger wastegate. _______________________________ 289 © Cummins Turbo Technologies. martensitic and sintered powder metal alloy steels. Since the interface of interest resides within the turbine housing of the turbocharger. Wear models for the various material combinations have been proposed and related to those experienced by the wastegate whilst in operation. cobalt-chromium alloys and a wide range of austenitic. which was operating at 600 °C. namely. The wear performance of the evaluated material combinations have been determined and related to the topography and chemistry of the new and worn surfaces in order to ascertain the fundamental wear behaviour of the alloys under investigation. in order to afford sufficient wear behaviour and performance for the interface. A wide range of shaft and bush materials were evaluated for their suitability within the previously mentioned interface. temperatures exceeding 600 °C are experienced by the tribosystem. 2014 . bushings and shafts in modern turbomachinery wastegates have been produced from stainless steel or cobalt-based alloys. it is imperative that the material combination for the shaft / bush tribosystem is extensively characterised in order to afford sufficient confidence while in operation. 3). INTRODUCTION Turbocharging is a fundamental technique employed by engine manufacturers in order to minimise emissions. Traditionally. Cummins Turbo Technologies. UK b Advanced Engineering. Cummins Turbo Technologies. Materials selected for use within the interface had to possess adequate friction and wear performance whilst in operation. maximise power and reduce fuel consumption of the internal combustion (IC) engine. it is important to characterise the tribological performance of particular components. In order to ensure high levels of reliability and durability of the turbocharger. 1. Contact pressure. However. Therefore. The latter material is selected due to excellent strength. the maximum exhaust gas temperature experienced by the turbocharger turbine housing can vary significantly. corrosion and tribological performance (1) over a wide temperature range.The high temperature tribological performance of turbocharger wastegate materials M Burkinshawa. sliding distance and speed were kept constant throughout testing. Since turbocharging is used in conjunction with diesel. the price of such materials is prohibitive for employment within cost-sensitive applications and therefore. 850 °C and 950 °C. alternative lower cost alloys are under consideration (2. D Blackerb a Materials Engineering. itself a fundamental component within modern turbomachines. UK ABSTRACT The high temperature tribological interface between shaft and bush in a turbocharger wastegate has been investigated using a pin-on-plate reciprocating tribometer. gasoline and natural gas engines. 5 ± 0. Secondary Ion Mass Spectroscopy (SIMS) was used to determine the tribochemistry of worn substrates. including the valve. EXPERIMENTAL 2.6 ± 0. The latter is fixed in position within the turbine housing of the turbocharger and the shaft rotates within the bush about a given arc length. shaft and bush. The turbocharger wastegate shaft / bush interface Figure 1 is a computer aided design model of a typical turbocharger wastegate. effective simulation techniques will improve the efficiency and reduce the cost of product development. Table 1: Test substrates Material Pin / Disc Hardness (HRC) Combination 1 (austenitic stainless steel) / 2 A 18. Such testing is vital in order to validate materials and test conditions prior to more complex and expensive component-based experiments. High magnification topographical images of substrates were obtained using Scanning Electron Microscopy (SEM).1 / 23. 2. Ultimately.3 (intermetallic stainless steel) 7 (nickel superalloy) / 8 (sintered D 33. martensitic and sintered powder metal alloy steels have been presented.2 / 24.2.3 ± 1.4 powder metal stainless steel) 290 . The high temperature tribological performance of a range of austenitic.9 ± 0.5 / 51. Wear behaviour and results are compared and related to those which would be experienced by a wastegate in operation. together with their cost and predicted tribological and corrosion performance. Figure 1: Typical wastegate shaft / bush 2.3 ± 0.1. these are listed in Table 1.5 ± 0.The intention of this article is to present the innovative methods by which conventional and novel materials intended for use within the wastegate system of a turbocharger are characterised using a tribometer and associated surface characterisation techniques.4 5 (austenitic stainless steel) / 6 C 20. Tribological testing affords the acquisition of accurate and repeatable friction and wear data from substrates of interest.9 ± 0.2 (martensitic stainless steel) B 3 (cobalt alloy) / 4 (cobalt alloy) 38. whereas combinations A & B have found previous use on turbochargers with a turbine inlet temperature < 800 °C. The test materials were selected due to their suitability for a wastegate application.9 ± 0. Material combinations C and D were novel selections in order to cater for high temperature wastegate applications. Materials Numerous high temperature materials were evaluated.3 / 52. ν2 = Poisson’s ratio bush & shaft P = Applied force (N) E1. which afforded sufficient comparison between the tribological performance of material pairs.2 µm Ra. with a surface roughness value of 0. it thereby induces a top and bottom dead centre onto the wear scar.5 = πa + Equation 1 (4) Equation 3 (4) = 0. P (σ ) = 1. The effect of exhaust gas pulsation was considered negligible for the purpose of the testing reported here. although with careful consideration of external factors which may affect the correlation performance between tribometer and real-world tests. Pins representing the wastegate shaft were manufactured with dimensions 10. 10.0 mm depth. this type of interface can be successfully simulated using a reciprocating pin-on-plate/disc tribosystem.721 √( ) 1− 1− Equation 2 (4) = + Equation 4 (4) σc = Contact stress (Pa) ν1. 291 . All substrates were used in as-received form without heat treatment.0 mm diameter.Since the turbocharger wastegate shaft / bush interface operates through rotary motion over a short arc length. with P determined using a free body diagram of the reaction force of the shaft in the bush. The peak contact stress for the shaft / bush interface was calculated using equation Equation 1-4 (4). Disc samples simulating the wastegate bush were machined with 35. The tribometer frequency of reciprocation and time duration of test were established by assuming the wastegate performed 216 x 103 cycles. 2.0 mm length and 37.0 mm diameter. The contact conditions (Table 2) for the tribometer were designed to replicate those experienced by Figure 2: Schematic diagram of pin- a wastegate in a Cummins Turbo on-reciprocating plate tribometer Technologies D6 turbocharger. 30.3. Therefore.5 mm radius of curvature on the end of the pin. Methods Tribological evaluation of the test materials was conducted using a modified version of ASTM G133 and a Bruker UMT-3 high temperature pin-on-reciprocating plate / disc tribometer (Figure 2). E2 = Young’s modulus bush & a = Contact radius (m) shaft (Pa) Kd = Equivalent diameter (m) D1 = Bush inner diameter (m) CE = Equivalent elastic modulus (Pa) D2 = Shaft outer diameter (m) The relative bush / shaft displacement was determined by considering a typical angular rotation of the wastegate shaft of 15 degrees and a shaft diameter of 10 mm. The previously mentioned number of cycles would equate to 87 x 103 km on engine. Samples were submerged in acetone and ultrasonically cleaned for 10 minutes prior to each test. namely materials 1.2. Table 2 states the test parameters for the tribological experiments. 292 . Table 2: Tribological test parameters Parameter Value Maximum contact pressure 180 MPa Total stroke length 1. The tribochemistry of the worn substrates was obtained using a ToF-SIMS IV Secondary Ion Mass Spectrometer. 3. Furthermore.The maximum predicted contact stress for the shaft / bush interface was replicated for the pin-on-plate tribosystem using an applied load of 4 N. Volumetric wear data was obtained from three repeats per material combination using an Alicona™ InfiniteFocus® confocal microscope. 5 and 7. 850 °C and 950 °C. 850 °C and 950 °C Test duration 2 Hours Test Medium Air 2. Low magnification images were obtained using a Leica® DM LM microscope. The wear resistance of combination D was superior to combination C at 600 °C. RESULTS 3. it was not possible to accurately measure the wear volume for this material combination at 950 °C because of an excessive oxide formation rate. except at 600 °C for the pin. which had a lateral and vertical resolution of ≤ 2.3 mm Frequency 30 Hz Operating temperatures 600 °C. 600 °C The wear scar induced into the substrates tested at 600 °C are shown in Figure 4.4.2 μm and 410 nm. Volumetric wear data Combination A possessed the worst wear resistance at all temperatures evaluated. The cobalt-based material combination B possessed the greatest wear resistance of all four material pairs at all testing temperatures. Figure 3: Wear volumes measured on (A) pins and (B) discs 3.2. appeared to consist of numerous abrasive wear tracks which were of non-homogenous height distribution. 3. respectively. The worn regions of all pin samples.1. Images of worn substrates 3.1. Evaluation of worn substrates High magnification images of surfaces of interest were obtained using a Philips™ XL30 Scanning Electron Microscope. 2. Arrow indicates sliding direction. 4. at 850 °C. The wear scar induced into material 5 was dominated by abrasive wear tracks.3. Figure 4: Experiments conducted at 600 °C. There were regions within the wear scars of materials 6 and 8 which appeared to be indicative of material removal. 4. 5 & 7) and 200x magnification SEM images of the wear scar on test plates (materials 2. The worn regions in materials 3 and 4 possessed a smooth appearance. 3. as shown in Figure 6. 3. 6 & 8). 293 . The wear scar of materials 6. 6 & 8). Arrow indicates sliding direction. 5 & 7) and 200x magnification SEM images of the wear scar on test plates (materials 2. The wear scar of material 6 appeared uneven and contained a number of pores. In addition. In contrast. 6 & 8. 4.2. but were also non-uniform in height distribution. 850°C As can be observed in Figure 5. Figure 5: Experiments conducted at 850 °C. there was a definable lay to the wear scar of material 8. the wear scars observed on materials 4 and 8 appeared smooth. in which a topographically smooth surface was generated post-test. 950°C Non-uniform dimensioned oxide particles were observed in the wear scar of material 2 when operating at 950 °C. This was in stark contrast to materials 3. 3. 7 and 8 were relatively smooth. whereas the wear region of material 7 was smooth but contained numerous pores.The wear scar generated in material 2 appeared extremely course and multiple abrasive wear tracks were evident throughout. 20x magnification optical microscope images of the wear scar on test pins (materials 1. 20x magnification optical microscope images of the wear scar on test pins (materials 1. Multiple abrasive wear tracks were evident in materials 1 and 5. the wear scars generated in materials 1 and 2 were uneven and appeared to contain wear debris throughout. with the latter containing a number of raised sections inside the wear scar.2. 3. 3. Figure 6: Experiments conducted at 950 °C. Ion intensity is represented by the scale on the images. Chemical nature of worn substrates 3. Arrow indicates sliding direction. Figure 7 shows a typical series of images obtained from the cobalt-based material pair. 6 & 8). 4.3. 294 . 20x magnification optical microscope images of the wear scar on test pins (materials 1. A & B are obtained from material 3 (pin) and C & D from material 4 (disc). The main elements and oxides identified within the wear scar on both pin and disc are stated in Table 3. combination B. Total ion count is listed as “TC”. 3. This technique afforded direct comparison of the chemistry of regions inside and outside the wear scar. Figure 7: Typical chemical images for Co and Cr obtained using ToF-SIMS. 5 & 7) and 200x magnification SEM images of the wear scar on test plates (materials 2.1. 3. SIMS Chemical images of worn substrates were obtained using secondary ion mass spectrometry. 3. D 8 (sintered powder CrO2. further work will be required to define whether the observed material transfer reduced wear in the interface through steel / steel rather than steel / nickel alloy interaction. FeO2 CrO2. MnO2 steel) 7 (nickel superalloy) / Cr. Cr. O. which were in the order of 200 MPa. MnO2 Co. Disc – Material Pin / Disc Elements of Elements of Combination Interest Interest 1 (austenitic stainless Cr. Fe. Interestingly. Mo. CrO2. CrO2. Fe. 3 (cobalt alloy) / 4 Co. stainless steel) FeO2. Referring to section 3. Stainless steel pair (combination A) The wear protective afforded by cobalt-based alloys is well reported in literature (1. CrO2. Mn. 3). Cr. Fe. O. with material combination A possessing the lowest yield strength of all substrates evaluated at 90 MPa. effective high temperature wear materials must retain strength and hardness when at temperature and the results presented here agree with this statement. Fe. from best to worst performing: 1. MnO2 5 (austenitic stainless Cr. MnO2 FeO2. Table 3: Chemical composition of wear scars Pin . O. Ni. Stainless steel pair (combination C) 4. (intermetallic stainless FeO2. Mn. O. However. In contrast. it is apparent that the wear experienced by material combinations A. As stated by Inman et al. B and C are predominantly of the abrasive type. DISCUSSION 4. it is apparent that the materials with the greatest yield strength (Table 4) possessed the greatest wear resistance. B O. Mn.. Nickel super alloy / stainless steel ( combination D) 3. A steel) / 2 (martensitic Ni. Ni. Mn. 5) and it has been reported that nickel-based alloys possess inferior wear protection compared to their cobalt-based counterparts (6. (7). Tribological performance of test substrates Referring to section 3 of this article. 295 . (cobalt alloy) CrO2. O. Indeed. It can therefore be inferred that such behaviour influenced the wear performance afforded by this material combination. steel) / 6 C O. Cobalt-based super alloy pair (combination B) 2. it is hypothesised that adhesion also contributed to the wear behaviour in this interface. Indeed. due to the evidence of material transfer with material combination D.1. it is apparent that the wear performance of the material pairs evaluated in this article at both 850 °C and 950 °C can be classified in the following order. Fe. Fe. CoO CoO. the high concentration of Fe and FeO2 on the nickel- based pin is evidence that material transfer from disc to pin occurred with material combination D. CrO2. FeO2 metal alloy steel) 4. CrO2. O. Cr.1.. Cr. MnO2 FeO2. referring to section 3. the stainless steel material pair C possessed considerably lower yield strengths at 850 °C. (10). Referring to the condition of material combination A post test. in order to provide a reliable and accurate baseline by which comparisons could be made to tribometer test samples. on the contacting surfaces. Referring to the topography of the glaze generated on the cobalt based material in Figure 5 and that reported on numerous materials (10. An additional factor that contributes to the wear resistance is glaze formation and the chemistry of the oxide layer which is considered in the following section. 11. Tribometer testing as a simulation for wastegate wear behaviour On the intended engine application. Referring to Wood et al. However. it is apparent wear resistance was not dependent entirely on strength and additional factors may be contributing to wear performance. Table 4: Material yield strength at 850 °C Material Yield Strength (MPa) Cobalt alloy (3) 270 Cobalt alloy (4) 620 Nickel superalloy 680 Sintered powder metal alloy steel 400 The slight deviation from the rule is material pair D. 4. normal load. 9. It is clear therefore that both high strength and oxide layer stability are desirable characteristics of a high temperature wear resistant material pair. and operating temperature are fundamental parameters with regards to glaze formation for a particular alloy (9).. only provided the second greatest wear response. (10) it would appear that the wear scar generated on materials 3 and 4 shown in Figure 5 and Figure 6 closely resemble glaze formation on similar cobalt based alloys. Frequency of reciprocation. It was noted by the current authors that all other material pairs generated a more stable and smooth oxide layer post-test at all testing temperatures. Further work is required to determine the chemistry of the glaze through depth profiling. Therefore. 4. 10) and subsequently improves tribological performance of a given interface (8. 12). One such factor is the formation and stability of generated oxide layers.2. Furthermore. wastegate shaft and bush components produced from material combination A were sourced.3. Therefore. it is imperative that material combinations are selected for their ability to generate successful glaze layers which ensure a sustainable wear rate and frictional response in the tribosystem over the useful life of the component of interest. the wastegate shaft / bush will experience operating temperatures of up to 950 °C. a composite layer generated from oxidised wear debris (8). oxidation and mechanical mixing (8. the glaze generated on the cobalt- based materials was chemically similar to that identified by Wood et al. 10). it is suggested by the current authors that glaze formation has also occurred on materials 6 and 8 at 850 °C and 950 °C. Glaze formation and the effect on wear performance The ideal scenario for a high temperature tribological contact would be the generation of a glaze. compared to material combination A. which despite possessing the greatest strength. the oxide layer was topographically rough and became increasingly unstable with increasing temperature. These substrates were retrieved from a Cummins Turbo Technologies 296 . Glaze formation occurs through a complicated process of wear. However. it is hypothesised that excessive oxide formation. Therefore. Such additional information will increase confidence with regards to tribometer-based testing and provide further evidence that simulated wastegate testing can accurately and reliably determine the tribological performance of material combinations for a given wastegate application. The adhesive wear behaviour of the latter combination needs to be further investigated to ensure that this does not impact the durability of components whilst in operation. There was a strong correlation of material yield strength to 297 . since the former testing technique may permit a greater rate of debris removal compared to the latter. Superior wear performance was afforded by combination B. which was compiled of two cobalt-based alloys.turbocharger which had operated on engine with a maximum operating temperature of approximately 650 °C. could potentially reduce the clearance between shaft and bush whilst in operation. Variation in wear debris behaviour may affect tribochemistry. 850 °C and 950 °C. which were performed at 600 °C. Abrasive wear was the dominant wear mechanism in the field return shaft / bush tribosystem. All material combinations evaluated in this research provided significantly greater wear resistance than material combination A at all testing temperatures. CONCLUSIONS A wastegate shaft / bush interface has been successfully simulated using tribometer-based experiments. this is justification to suggest that the research and methodology reported here provides a cheap. 5. The current authors are of the opinion that tribometer-based testing can remove the need for particular functional tests. in terms of wear rate and tribochemistry on both substrates. high-efficiency tool to characterise high temperature tribological materials for wastegate applications and identify wear mechanisms and oxide / glaze layer formation. as experienced with material combination A at 950 °C. glaze formation and wear protection. One item of interest for the current authors is the behaviour of wear debris produced through sliding in both the pin-on-disc and shaft / bush interfaces. Figure 8: Worn shaft and bush from field-return turbocharger Further testing and analysis is required in order to provide additional data points for correlation. It is anticipated that further research into this topic will improve the correlation of simulated and application testing. improving the efficiency and reducing the cost of product development. more stable alloys such as combinations B & D should not suffer such issues. which correlated well to the results presented for material combination A. Contact conditions were derived from a production turbocharger. Referring to the wear behaviour of the four material combinations. S R and Datta. L. Wear. pp. Vol. 2006. Vol. W C and Budynas. Inman. 4. 2008. 461-467. Liu. Inman. Bodies under Direct Bearing and Shear Stress. 38. Microscopy of Glazed Layers Formed during High Temperature Sliding Wear at 750 C. H. 2. pp. Additional work is also required to generate further evidence for the correlation accuracy and reliability between tribometer- based and application testing of wastegate materials. Wear. R and Wu. 812-823. 263-271. 10. Roark's Formulas for Stress and Strain. Y. Investigation into the Wear Behaviour of Tribaloy 400C During Rotation as an Unlubricated Bearing at 600 oC. The Galling Wear Resistance of New Iron-Base Hardfacing Alloys: A Comparison with Established Cobalt. 12. 8. Vol. Improved Mechanical and Tribological Properties of Tin-Bronze Journal Bearing Materials with Newly Developed Tribaloy Alloy Additive. 2002. Wear. Materials Science and Engineering A. I A. Vol. 389- 402. 76-77. 702-703. R G. pp. pp. B. 265. R J. Galling Mechanisms during Interaction of Tool Steel and Al-Si Coated Ultra-High Strength Steel at Elevated Temperature. Hardell. 13. Wear. 11. 9. Improvement of the Oxidation Resistance of Tribaloy T-800 alloy by the Additions of Yttrium and Aluminium. Vol. Ocken. Roark. pp. Vols. pp. Vol. Further work is envisaged in order to understand the effect of wear debris on glaze formation and tribological performance. 664-671. Studies of High Temperature Sliding Wear of Metallic Dissimilar Interfaces. 202. 489. 456-461. H E and Ponton.tribological performance. 7. 5. pp. 1995. Vol. P S. Characteristics of Tribaloy T-800 and T-900 Coatings on Steel Substrates by Laser Cladding. P D. Vol. Surface and Coatings Technology. 2005. 1819-1827. 2297-2301. Vol. 67. : McGraw Hill. Zhang. 269. et al. 254. with further wear protection afforded by the generation of stable glaze layers within the wear scars on tested substrates. pp. 269. Birol. 6. 39. Rose. 7. 298 . Room Temperature Mechanical Properties and Tribology of NICRALC and Stellite Casting Alloys. High Temperature Sliding Wear Behaviour of Inconel 617 and Stellite 6 Alloys. 271. Tribology International. W S. Inman. 6. et al. Material transfer from steel to nickel-based alloy was observed with material combination D. Pelcastre. 1592-1605. 2013. Tavakoli. Studies of High Temperature Sliding Wear of Metallic Dissimilar Interfaces II: Incoloy MA956 versus Stellite 6. et al. pp. P K. et al. Vol. Tribology International. X J. Development of a Simple Temperature Versus Sliding Speed Wear Map for the Sliding Wear Behaviour of Dissimilar Metallic Interfaces II. Inman. Tobar. pp. 1035-1043.and Nickel-Base Alloys. 763-769. Wear. M J. I A and Datta. Wood. Y D. I A. 2011. 1361-1375. Tribology International. pp. 2008. Vol. Vol. 2008. 3. Young. s. Corrosion Science. A. pp. I A. et al. pp. REFERENCE LIST 1. Evans. da Silva. 7. 2010. ACKNOWLEDGEMENTS Further thanks are given to David Scurr from The University of Nottingham for obtaining the SIMS data and Leigh Fleming from Huddersfield University for volumetric wear analysis. 53. C B. J and Prakash.l. Among such technologies. A standard turbocharger uses a single scroll turbine. An engine with a turbocharger. a radial turbine used for a turbocharger operates under exhaust gas pulsation. resulting in improvement of fuel efficiency and reduction of CO2. 2014 471 . turbochargers have been increasingly applied for downsizing of engines. Two features of flow phenomena generating loss were confirmed. Therefore use of turbochargers has been growing. Ltd. New scroll was designed to reduce alternating leakage inflow and flow separation at blade hub. On the other hand. can use smaller displacement than normal aspiration engines. _______________________________________ © The author(s) and/or their employer(s). Twin-entry scroll turbine gains increased attention because of high efficiency at low engine speed by utilizing engine exhaust gas pulsation. 1.Development of twin-entry scroll radial turbine for automotive turbochargers using unsteady numerical simulation T Yokoyama. and then the compressor bladed wheel. In addition. This corresponds to improvement of cycle average turbine efficiency. Japan ABSTRACT In recent years. regulations regarding emission gas and fuel economy of automobiles have been tightening year by year. T Yoshida. K Wakashima Mitsubishi Heavy Industries. Therefore turbochargers can reduce weight and friction loss of the engines. T Hoshi. which results in growth of the turbocharger market. An automotive turbocharger uses exhaust gas to rotate the turbine. at lower engine speed. it cannot compress air supply sufficiently because exhaust gas flow is small and the boost pressure is low. In particular European automotive manufacturers have been increasingly employing smaller engines for improvement of fuel efficiency. The authors confirmed partial admission efficiency at rear side is increased by 2 point via experiment. responding to growing interest in environmental issues. the shaft of which is connected directly with the shaft of the turbine. Numerical fluid simulations of twin-entry-scroll with whole turbine blade were conducted to investigate the flow phenomena under engine exhaust gas pulsation. Because large pulsation increases flow unsteadiness. Exhaust gas is induced into the front and rear scrolls alternately and leakage inflow from one side to no-entry side is occurred. In this study. in particular many automotive technologies for fuel efficient vehicles and countermeasure to emission control have been developed actively. is driven. However. a twin scroll turbine is effective(1). To raise boost pressure. Flow separation at blade hub side is occurred by the distorted inflow from the scroll at the partial admission of rear side. which feeds compressed air to the engine. it was difficult to understand flow phenomena and improve turbine efficiency. INTRODUCTION Along with growing interest in global environmental issues. For further enhancement of fuel efficiency it is necessary to improve efficiency of turbochargers. Fuel consumption at actual engine operating conditions such as low engine speed are focused. This document presents work on improvement of twin scroll turbine efficiency with use of the unsteady computational simulation. because internal flow of a twin scroll turbine was made unsteady by largely pulsing flow. Fig. and also its detail has not been well understood. MHI. NUMERICAL SIMULATION METHOD AND SIMULATION CONDITIONS In the past. Improvement of fuel efficiency at lower engine speed requires enhancement of turbocharger efficiency. has taken its advantage to perform unsteady analysis that simulates exhaust pulsation as shown in Figure 1 (ii). NEDC (New European Driving Cycle)(3) has been used since its introduction in 2000. unsteadiness of internal flow that actually occurred as a result of exhaust pulsation flowing into the two scrolls alternatively and intermittently was unknown. 3. Because such conventional analysis used a constant pressure that represented a mean exhaust pulsation condition. Now European countries are planning test cycles where fuel economy and emission gas of vehicles are measured under conditions simulating actual driving conditions as closely as possible. To analyze unsteady phenomena under conditions of exhaust pulsation. 1 Turbine pressure Fig. Therefore it will be necessary for complying with such regulations to improve fuel efficiency at lower engine speed further than ever. in the future. The analysis objects were entire turbine including the scroll and the turbine rotor blade (turbine diameter 52mm.3. They are considering use of CADA (Common Artemis Driving Cycle)(4) etc. it is difficult to analyze the internal flow. it is predicted that a driving mode with repeating acceleration from lower engine speed and deceleration will be added in consideration of driving in urban area. The numerical flow analysis used ANSYS CFX general-purpose 3D viscous flow simulation code. a twin scroll turbine was analyzed under steady operating conditions as shown in Figure 1 (i).07 million cells on the turbine rotor blade and 1. blade number 12). However. The turbulent model used was the k-ε model. As shown in Figure 2.Fuel economy shown on automobile brochures is measured according to a certain driving pattern (test cycle)(2) designated in the country or local region.44 million cells on the scroll. This analysis was performed at lower engine speed of 2400 rpm (turbocharger speed of 125700 rpm). 2. Figure 6 (i) illustrates internal flow results of conventional steady analysis at a constant pressure as will be described in 3. In such future test cycles. the computational grid used contained 2. MHI performed flow analysis using CFD (Computational Fluid Dynamics). 2 Grid for analysis 472 . In European countries. which is a manufacturer of both engines and turbochargers.51 million cells in total. a little from three-cylinder.1 Analysis of internal flow phenomena of turbine under conditions of exhaust pulsation For a conventional shape. Therefore the most important factor for performance evaluation of a twin scroll is when inflow from one side occurs. The generated loss flows down in circumferential direction of the scroll. 3 Exhaust pulsation in twin scroll (combined with four-cylinder engine) 3. Figure 3 illustrates exhaust pulsation in a twin scroll combined with a four- cylinder engine (Engine simulation results at turbine inlet). MHI checked internal flow on the rear and front cross sections. and propagates through the rotor blade to the outlet. in-leakage to the 473 . It is assumed that exhaust efficiency of a single scroll combined with a four-cylinder engine is reduced by this interference. which is not fed.There occurs no exhaust pulsation interference from a two. On the other hand. It is assumed that because a twin scroll. generates pressure difference between each scroll. Figure 4 illustrates distribution of loss (i. Figure 5 shows loss distribution and flow distribution on cross sections of the rear side where main stream is being fed and the front side where no stream is being fed during feeding of the rear side (Figure 1 (ii) (a)). entropy) inside a turbine during inflow of exhaust gas from the rear side. where exhaust pulsation flows into the front and rear scrolls intermittently.or one-cylinder engine. NUMERICAL SIMULATION RESULTS 3. and intense from four-cylinder. Fig. As time passes. each scroll of which corresponds to a scroll connected to a two-cylinder engine. To analyze the cause of increase of loss in the front side scroll not being fed during inflow from the rear side. Loss is generated in the scroll with time. becomes larger. loss in the front side scroll.e. It is found that there is leakage in the non-fed side which caused large loss. To understand the cause of lowering of turbine efficiency during inflow from the rear side. it was found that mean in-cycle turbine efficiency during inflow from the rear side was lower than that during inflow from the front side. Analysis of operation of a twin scroll shows that partial inflow from one side occurs in almost all range of engine crank angle and perfect full inflow from both sides occurs only at points where pulsation switches. MHI focused attention on formation of internal flow loss during inflow from the rear side. there occurs no exhaust interference for a twin scroll. 4 Internal flow of turbine during inflow from rear side (a) Fig. 5 Internal flow of turbine scroll during inflow from rear side (a) 474 . and then enters into the rotor blade. making loss and causing efficiency lowering. The in-leakage flow into the lower pressure side flows in the scroll for a while. Fig.lower pressure scroll occurs. Unlike inflow from the rear side. flow separation this time is smaller than that during inflow from the rear side. MHI has focused attention on flow separation inside the rotor blade and leakage flow to the non-fed side in the scroll to make improvements for a new twin scroll turbine.Figure 6 (ii) (a) shows internal flow of the turbine on a cross section at a constant position in the circumferential direction during inflow from the rear side (Figure 1 (ii) (a)). Figure 6 (ii) (b) shows internal flow of the turbine on a cross section at a constant position in the circumferential direction during inflow from the front side (Figure 1 (ii) (b)). 6 Internal flow of turbine on a cross section at a constant position in the circumferential direction 3. Fig. Figure 7 compares leakage flow in the non-fed side during inflow from the rear side (Figure 1 (ii) (a)). and then flow separation occurs on the shroud side inside the rotor blade. but it is found through unsteady analysis that exhaust pulsation flowing into the front and rear scrolls alternatively and intermittently causes flow separation loss in the rotor blade and this loss leads to efficiency lowering as transferred to the rotor blade outlet. However. Conventional analysis under inflow from both sides at a constant mean pressure of exhaust pulsation could not find out internal flow loss. no flow separation is found on the hub side inside the rotor blade during inflow from the front side. Leakage flow from the throat of the front side scroll flows to the hub side of rotor blade. 475 . It is found that leakage flow from the throat of the rear side scroll flows to the shroud side of rotor blade. and then flow separation occurs on the hub side inside the rotor blade.2 Improvement focusing attention on flow separation inside rotor blade and leakage flow in scroll Among internal flow phenomena described above. On the other hand. As shown in this comparison. flow distribution of the new twin scroll has a smaller low flow rate area in the hub side than that of the conventional twin scroll. loss distribution on the 90% span cross section of the new twin scroll has a smaller loss at the posterior border of the shroud than that of the conventional twin scroll. the new twin scroll has achieved 1. Figure 9 compares turbine efficiency. which results in suppression of flow separation during inflow from the rear side. Loss at the trailing edge of the shroud is originated in loss generated in the rotor blade hub side. 476 . As shown in Figure 8 (ii).The conventional twin scroll turbine had leakage flow at the position around 90 degrees in the circumferential direction. Figure 8 compares internal flow of the turbine rotor blade between conventional and twin scrolls during feeding of the rear side (Figure 1 (ii) (a)). Fig. 7 Comparison of leakage flow on cross section in front side between conventional and new twin scrolls during inflow from rear side (a) Next. Therefore this reduction of loss means improvement of internal separation flow of the rotor blade. As shown in Figure 8 (i). but the new twin scroll can achieve its reduction and improvement of flow phenomena.9% enhancement of mean in-cycle turbine efficiency mainly due to improvement of efficiency during inflow from the rear side. MEASUREMENTS OF TURBINE PERFORMANCE Performance tests of new twin-scroll turbine were conducted on gas-stand (Fig. Fig. Turbine diameter is 43mm and blade number is 11. 10). turbine performance is improved at rear flow condition by 2%. This improvement will enhance pulsation 477 . Efficiency at Front flow and full flow condition are slightly reduced. 8 Comparison of internal flow of turbine rotor blade between conventional and new twin scrolls during inflow from rear side (a) Fig. On the other hand. 9 Comparison of turbine performance between conventional and new twin scrolls 4. php 478 . 2012 http://www.dieselnet. Also MHI will promote optimization of the shape of a twin scroll turbine for use with a small engine.5 2. Emission Test Cycles ECE +EUDC/NEDC.0 2. using unsteady computational simulation for efficiency enhancement of turbocharger. 2000 http://www. and demand for higher efficiency of turbochargers seems to be endless. 6..5 2.com/standards/cycles/ece_eudc.dieselnet. taking into account its productivity.0 Turbine pressure ratio [-] Turbine pressure ratio [-] (a) Partial admission (b) Full admission Fig. MHI is willing to develop higher-efficiency turbochargers to improvement of environmental performance in the future. This twin scroll turbine can raise boost pressure even at lower engine speed. and therefore improve fuel efficiency.0 2. FUTURE ISSUES AND OUTLOOK MHI has obtained prospects of improving turbine efficiency through restraining internal flow loss of the twin scroll turbine under conditions of exhaust pulsation.5 3.0 1. This characteristics also reduce engine performance deviation between cylinders. Summary of worldwide engine and vehicle test cycles. Emission Test Cycles Common Artemis Driving Cycles (CADC). 10 Gas-stand test results 5. MHI will conduct performance tests on an engine test bench to verify the effect. Mitsubishi Heavy Industries Technical Review Vol. In the future. It is predicted that tightening of regulations regarding emission gas and fuel economy of engines will continue in the future. CONCLUSION Responding to tightening of regulations regarding emission gas and fuel economy of automobiles. M. et al.dieselnet.1 (2004-1) (2) DieselNet.php (4) DieselNet. Then deviation of turbine performance between front flow condition and rear flow condition is reduced for new twin scroll turbine. 2011 http://www.com/standards/cycles/artemis. Mitsubishi Turbocharger for Lower Pollution Cars. it is necessary to improve accuracy of simulation by measuring the unsteady experimental data that verifies unsteady computational simulation. Because internal flow of the turbine under conditions of exhaust pulsation is highly complicated. increase engine torque. Base-Fside Modified-Fside Base-All Modified-All Base-Rside Modified-Rside Turbine efficiency [-] Turbine efficiency [-] 5% 5% 1.0 1.com/standards/cycles/#eu (3) DieselNet.0 1.cycle average efficiency of turbine.5 3. 41 No. REFERENCES (1) Ebisu. 7. MHI has developed a high-efficiency twin scroll turbine. ............................................. 149 Burkinshaw........................................ AUTHOR INDEX Aghaali...... M V ................. T .............................. R .......................... Z .................. 179 Akehurst.......... 289 Bolz............ L ........................................................ 13........................ 149 Duda........................... E ............. 89 Dowell.............. 361 Baar............................... A ................................. 301 Binder........ 241 Criddle. 149 Casey............................................................. 27...................... 113 Christen............................... 149........................... A . 27 Cornwell....... 289 Capon... 79 Andrews......................... 149 ............................. 449 Brace.................................. P ................................... M .................................... H ................. C ......... 137.............. C J ........................................ 41 Bou-Saïd............ 3 Banks...... G .... D ... M ......................................... 163 Alsalihi.......................................................... H .............. 163 Burke.................... 13...... B ...... C .......................... 103 Ashtekar............................. 149 Dietrich.............. 55 Chen..... 27........................ R .................................................................................. 253 Ångström.................... H E ....................... D N ............................. 13....................... 41 Blacker....................................................... F J .... 149... 241 Bargende......... P ............. 103......................................... H ................................................. 179 Arnau........ E .................. B .......................................... M ........................... M .......................................................... 265 Böttcher.................................................................. 189 Copeland.. 189 Codan.. R D ..................... S ....................... 207 Davies....... 65 An.......................................................... Early, J .............................................................. 89 Eynon, P ........................................................... 399 Filsinger, D .......................................... 89, 301, 349 Friedrich, I ........................................................... 3 Garrett, S ......................................................... 163 Grabowska, D .................................................... 227 Grigoriadis, P ..................................................... 41 Groves, C .......................................................... 137 Gurunathan, B A ................................................. 13 Hagemann, T .................................................... 375 Harley, P X L ...................................................... 89 Hattori, H .......................................................... 389 Hazby, H R ......................................................... 55 Heyes, F ........................................................... 125 Hirai, Y ............................................................. 217 Hoffmann, R ...................................................... 437 Hoshi, T ............................................................ 471 House, T ........................................................... 227 Hu, B ................................................................ 27 Ibaraki, S ........................................................... 65 Ikeya, N ........................................................... 217 Jackson, R .................................................. 13, 207 Jarvis, S ........................................................... 207 Kadunic, S ........................................................... 3 Kasthuri Rangan, P S .......................................... 281 Kaufmann, A ..................................................... 265 Köhl, W ............................................................ 349 Kreschel, M ....................................................... 349 Lamquin, T ........................................................ 449 Lancaster, C ...................................................... 361 Lee, D .............................................................. 207 Lewis, A G J ....................................................... 13 Liebich, R .......................................................... 437 Liu, S T ............................................................. 411 Liu, Y H ............................................................ 411 Lotz, R .............................................................. 227 Luard, N ..................................................... 13, 207 Lüddecke, B ...................................................... 301 Marques, M ....................................................... 137 Martinez Botas, R F ............................... 13, 321, 333 Matteucci, L ....................................................... 13 Morand, N ......................................................... 137 Moscetti, J ........................................................ 227 Mrazek, R ......................................................... 137 Nanbu, T ............................................................ 79 Newman, P ....................................................... 207 Numakura, R ...................................................... 55 Okhuahesogie, O F ............................................. 125 Olmeda, P ......................................................... 103 Padzillah, M H .................................................... 333 Pandian, S ........................................................ 281 Parikh, S S ........................................................ 281 Porzig, D ........................................................... 421 Pronobis, T ........................................................ 437 Raetz, H ........................................................... 421 Rajoo, S ........................................................... 333 Ramamoorthy, J M ............................................. 281 Rémy, B ........................................................... 449 Reyes-Belmonte, M ............................................ 103 Richardson, S .................................................... 207 Riley, M J W ...................................................... 125 Rinaldi, A .......................................................... 265 Roach, P ........................................................... 125 Rochette, C ....................................................... 207 Romagnoli, A ............................................... 13, 321 Sakai, M ........................................................... 321 Sato, W ............................................................ 389 Scherer, F ............................................................ 3 Schwarz, J B ..................................................... 253 Schwarze, H .............................................. 375, 421 Scott, S ............................................................ 227 Sens, M ............................................................. 41 Seume, J R ....................................................... 421 Shi, X ............................................................... 411 Shoghi, K .......................................................... 461 Smith, L ........................................................... 163 Smith, T ........................................................... 207 Spence, S W T .................................................... 89 Stewart, J ......................................................... 125 Such, C ............................................................ 241 Sugimoto, K ....................................................... 65 Suzuki, T .......................................................... 217 Tamaki, H .......................................................... 55 Tang, H ............................................................ 163 Tian, L .............................................................. 361 Tomanec, F ....................................................... 137 Tomita, I ...................................................... 65, 79 Toussaint, L ...................................................... 137 Turner, J W G ................................................ 13, 27 Van den Braembussche, R .................................... 65 Vemula, R ......................................................... 227 Verstraete, T ...................................................... 65 Vetter, D ........................................................... 375 Vlachy, D .......................................................... 137 Wakashima, K ................................................... 471 Wan-Salim, W S-I ............................................... 13 Watson, J .......................................................... 227 Winterbone, D E ................................................ 113 Yamagata, A ..................................................... 389 Yang, C ............................................................ 411 Yang, M Y ......................................................... 333 Yokoyama, T ..................................................... 471 Yoshida, T ......................................................... 471 Zangeneh, M ..................................................... 399 Zatko, M ........................................................... 137 Zhang, J ........................................................... 399 Zhang, Q .......................................................... 149 Zhao, B ............................................................ 411 Zhuge, W L ....................................................... 333 Cool2Power - Increased petrol engine power and efficiency through an AC driven intercooling system S Kadunic 1, F Scherer 2, R Baar 1, I Friedrich 3 1 TU Berlin, Germany 2 Bundesanstalt für Materialprüfung und -Forschung (BAM), Germany 3 IAV GmbH, Germany ABSTRACT The supercharging potential of SI engines in passenger cars is mainly limited by self-ignition and turbo charger turbine temperature. A reduced charge air intake temperature would allow for a higher boost ratio and engine load or an increase in efficiency. The research in this paper shows experimental results using a turbocharged 1.4l SI-engine to prove the concept of extreme charge air cooling using a modified air condition system. Maximum torque was increased by up to 12.5%. Assuming the original full load limit, fuel consumption could be reduced by 9% in some operating points. 1 INTRODUCTION The downsizing of combustion engines has greatly improved the fuel efficiency of passenger vehicles in the common driving cycles like NEDC and FTP75. There is a trend to further decrease engine displacements to help fulfil future legislations regarding CO2-emissions and fuel consumption [1, 2]. At the same time it can be assumed that vehicle performance must not suffer for marketing and sales reasons. Under the constraints of passenger safety, comfort and cost, a reduction of vehicle mass will be limited. For those reasons, rated engine power most likely will not be reduced. As a fact, the average rated output of passenger vehicles sold in Germany increased by 42% from 1995 to 2012 [3]. To achieve the demanded power levels with downsized engines, superchargers (mostly turbochargers) are used and boost ratios are continuously increased. Unfortunately, turbocharged spark-ignition engines optimized for partial load operation show unfavourable properties at high load levels due to limits of engine operation. The occurrence of engine knock demands a late centre of heat release (COHR) [4]. This directly increases brake specific fuel consumption (BSFC) due to a decrease in thermal efficiency. Additionally, late combustion increases exhaust gas temperature. This leads to another restriction in the operation of these engines, which is the thermal limit of the turbocharger turbine [5]. Particularly at high engine loads and speeds, the exhaust gas temperature has to be decreased by running rich air/fuel-ratios. The evaporation enthalpy of the surplus fuel is employed to cool the turbine, which significantly decreases efficiency as this unburned fuel cannot contribute to power generation [6]. _______________________________________ © The author(s) and/or their employer(s), 2014 3 A decrease in intake air temperature reduces the knock tendency at a given engine load. This would permit earlier COHR which decreases BSFC directly and also reduces exhaust temperatures. Even at a given COHR, a reduction of intake temperature reduces the exhaust temperature at high engine loads. This effect can be employed to reduce mixture enrichment to increase efficiency at points of engine operation where turbocharger turbine temperature is of concern. As shown in [7] at the same time a significant reduction in exhaust emissions can be achieved. Regular production vehicles are equipped with intercoolers which can cool the intake air directly or indirectly via a liquid system with ambient air. Depending on driving situations and the design properties of the system the intake air temperature achieved is commonly between 5K and 40K above ambient. Further lowering of the charge temperature can be achieved by turbo cooling or early intake valve closing (“Miller Cycle”). Both methods require an increase in boost ratio to achieve reference load. However, this creates additional requirements for the turbocharger compressor and increases back pressure from the turbo charger turbine. This can lead to an increased residual gas content which is contradictive to the idea of avoiding engine knock. Also, costly refined supercharging systems with two stage turbochargers often become necessary. Air conditioning (AC) systems are standard equipment in passenger vehicles sold in the EU at the time of this publication (2013). Such a system contains most components required to create an ultra-low temperature intercooling system. Also, they are designed to cool down the passenger compartment quickly after start-up with the engine at idle. During regular driving there is abundant cooling capacity available. The compressors of these devices consume power from the combustion engine. Regardless, a benefit in engine output or fuel efficiency is expected (at least in certain areas of engine operation). In this study, the concept of low temperature charge air cooling is investigated employing a modified vehicle AC system with a 1.4litre turbocharged SI engine and vehicle simulation. 2 CONCEPT AND EXPERIMENTAL SETUP The goal of the conducted experiments was to test the concept of low temperature intercooling with limited efforts in parts and vehicle development. A low- temperature intercooling system should use production parts from automobiles as much as possible and operate under regular driving conditions while requiring minimum changes to the vehicle. Basic preliminary considerations were undertaken regarding the layout of the system. It was found to be energetically useful to conventionally cool the compressed intake air as far as possible after turbocharger compressor, before entering the low temperature cooling system. The setup shown in Fig. 1 was chosen. Compressed air from the turbo charger compressor passes a conventional intercooler (IC). For reproducibility, an air to water heat exchanger was used as sometimes found in regular production vehicles. The engine dyno environment provided sufficient fluid to cool the intake air to not more than 40°C at all operating points. This represents the performance of a well- designed conventional intercooling system in average driving conditions, regardless of whether an air-to-air or air-to-water system is assumed. The authors consider this a conservative assumption, as in real world driving intake temperatures can be well above this temperature level [8]. 4 Fig. 1: Test setup After the conventional intercooler, the precooled intake air passes the extra intercooler (ICe) for low temperature. The same type of air-liquid heat exchanger is used here as for the IC. It is fed with a low temperature fluid (water with antifreeze) of around -5°C to 5°C provided by the modified AC system. This setup permitted air temperatures of 10°C and lower after ICe, while avoiding excessive build-up of ice from condensed intake air water. The setup was designed in a matter for water condensing in the ICe to be directly provided into the intake runner with no opportunities for puddles to collect. After each low-temperature test, the system was operated without ICe-activation until regular operating temperatures were reached to relieve the intake runners from ice and water before further experiments or shutdown. The only modification of the regular engine-driven AC compressor regarded the refrigerant volume flow control. It was made accessible to the dyno system in order to control the cooling output. Refrigerant used was common R134a. Compressed refrigerant passes a condenser, which is liquid cooled by the dyno environment. This provides consistent boundary conditions, but in a passenger vehicle this would be the car’s air cooled AC condenser. A common thermal expansion valve is used to control evaporation temperature. The evaporator is of the refrigerant-to-liquid type as commonly used in commercial cooling systems. Here the low temperature fluid is cooled after taking up heat from the intake charge. An electric pump is employed to circulate the low temperature fluid through the evaporator and the ICe. All engine related parts like throttle body, intake and exhaust manifolds, turbocharger, catalytic converter, etc., were taken from a production engine without modification (except added measurement probes). The parameters of the test engine are shown in Tab. 1. It is a state-of-the-art production engine, common for European midsize vehicles at the time of this publication (see [1]). 5 Tab. 1: Technical data of test engine Test engine data Spark ignition, homogenous direct injection, Turbo charged 4 cylinders inline, DOHC 16 valve Displacement: 1390 cm³ Nominal output: 90 kW Peak torque: 200 Nm Nominal output from 5000 to 5500 min-1 3 EXPERIMENTAL RESULTS Three series of experiments were run. In each series, conventional engine operation with regular intercooling was compared to operation with the low- temperature intercooling device engaged. For the first series, the test engine was simulated in two vehicles representative for midsize vans and light utility vehicles. Points of engine operation were derived assuming those vehicles performing stationary high speed motorway driving, like common on the German Autobahn. During the second set of experiments, the test engine was operated at its maximum nominal effective load at various engine speeds. The goal of those two test runs was to find out if gains in thermal engine efficiency can offset the additional power required to run the modified AC system for low temperature intercooling. For the final tests, areas of engine operation were selected where an increase in output would be of high interest. This was considered the case at low engine speed (low-speed torque) and in the area of maximum power output. The goal was to spread the torque band further. Here, the authors investigated if the benefits from the cooler intake charge would permit a higher effective engine output, regardless of the power drawn to drive the AC compressor. After the tests performed under the operating conditions mentioned, the engine was disassembled to investigate potential damage from water. No abnormal wear or damage was found. 3.1 Autobahn In Tab. 2, the data of the sample vehicles are shown. In Fig. 2, resulting points of operation for the engine are visualized assuming typical driving situations in high speed traffic. In all situations, constant speed on even surfaces and no wind were assumed. Tab. 2: Vehicle data for Autobahn simulation Van Delivery Front area 2,56 m² 2,8 m² Drag coefficient 0,34 0,38 Gross vehicle weight 2165 kg 2180 kg 6 220 200 180 179 km⁄h Brake torque in Nm 191 km⁄h 160 160 km⁄h Delivery 140 140 km⁄h 160 km⁄h 120 100 120 km⁄h 140 km⁄h Van 80 120 km⁄h 60 500 1000 1500 2000 2500 3000 3500 4000 4500 5000 5500 6000 6500 Engine speed in min-1 Fig. 2: Autobahn driving situations in engine map for van and light delivery vehicle Obviously, the heavier and less aerodynamic delivery vehicle creates larger engine loads at given vehicle speeds. In Fig. 3 the results for engine operation with regular and low temperature intercooling are shown. In the case of the van, improvements from low temperature intercooling can only be achieved at top speed. At all other vehicle speeds, the engine does not reach or barely reaches its operating limits. The engine is operated at stochiometric air-fuel ratio and COHR is just slightly later than thermally ideal. The additional energy consumed by the AC compressor offsets any gains from earlier ignition timing. Only at top speed operation does low temperature intercooling show promising results. Here, the engine is operating in enriched mode with regular intercooling to thermally protect the turbocharger turbine. With cooler intake air, the mixture can be leaned out, which overcompensates the power needed to operate the modified AC system. An improvement in BSFC of approx. 2.5% was achieved. For the delivery vehicle, improvements in BSFC can be achieved at vehicle speeds of 160km/h upwards. Due to the higher loads induced by the heavier and more wind resistant vehicle, the engine is operating in enriched mode earlier than the case in the van application. At top speed, BSFC can be reduced by approx. 3.6%. Results from both vehicles show that serious improvements in BSFC can be achieved as soon as the engine is required to be operated with enriched air-fuel ratios for turbo charger protection with standard intercooling. In cases where colder intake air “only” permits slightly advanced ignition timing there is a great chance for any thermodynamic improvements to be overcompensated by power consumption of the AC compressor of the low temperature intercooling system. Essentially the heavier and bulkier the vehicle, the larger engine loads, which increases the fuel saving potential of the low-temperature intercooling system. Further down-speeding will enable benefits in BSFC at lower vehicle velocities. Also, it should be considered that acceleration, head wind or uphill driving will put the engine in enriched operation mode at lower vehicle velocities than simulated here. 7 However. the test engine was operated at its nominal full load throughout the rpm-range. The resulting improvement in thermal efficiency overcompensates the power consumed by the modified AC-system and an improvement in BSFC of up to 3. the reduced knock tendency and lower pressure levels permit earlier COHR (see Fig. Results are visualized in Fig. Fig.00 1. 4g shows that the engine has to be operated with very late COHR of up to 30° crank angle after top dead centre (ATDC) instead of the desired 4° to 8° ATDC to prevent engine knock.00 Lambda Lambda 0.95 120 km⁄h 140 km⁄h 120 km⁄h 0.2 Increase of brake efficiency at full load During this trial. 4. 4d.80 0.90 160 km⁄h 160 km⁄h 0. which provides a further 8 .05 1. resulting in lower exhaust temperatures too. The approach was to search for an improvement in fuel consumption at full load for all engine speeds through low-temperature intercooling. 3: Results during Autobahn driving 3. Additionally. 40 40 30 30 TAir in °C TAir in °C 20 20 Van Delivery 10 10 0 0 1500 1500 Boost pressure Boost pressure 1300 1300 in mbar in mbar 1100 1100 900 900 700 700 15 15 COHR in °CA COHR in °CA w/o AC Compressor 12 12 with AC Compressor 9 9 6 6 300 300 be in g/kWh be in g/kWh 280 280 260 260 240 240 1.80 3000 3500 4000 4500 5000 5500 3500 4000 4500 5000 5500 Engine speed in min-1 Engine speed in min-1 Fig. the engine operates with an enriched mixture to limit exhaust temperature for turbo charger turbine protection.85 km⁄ 0. at low engine speeds (up to approximately 2000rpms) the air- fuel ratio can be stoichiometric even with regular intercooling. 4g). At engine speeds over 2000rpm.05 1. As seen in Fig. 4h).8% can be achieved (see Fig. A reduction in charge temperatures lowers the temperature level during combustion.95 140 km⁄h 0.90 0.85 191 h 179 km⁄h 0. Employing the low-temperature intercooling system COHR can be advanced by up to 5° CA. Both effects combined allow reducing the mixture enrichment while keeping the same exhaust temperatures. Overall.00 Lambda Auxillary load AC 18.2 K 30 TAir in °C 190 20 180 w/o AC Compressor ∆T = 33.05 NMEP in bar 19. the air density increase from temperature reduction lowers the required boost pressure to achieve nominal full load at all engine speeds. which demands decent turbocharger operation at very high gas mass flow.0 0.3 Gain in low end torque and rated power Currently the common downsizing applications of turbocharged SI engines with single stage turbochargers suffer from two contradicting demands.0 Compressor 0.85 950 90 TExh. In points of operation with rich air-fuel ratio. 4: Results at nominal full load 3.95 17.0 1.reduction in the exhaust temperature. At the same time.90 c d λ<1 16. but low- temperature intercooling does not permit stoichiometric mixture at those points of engine operation. during nominal full load operation the low temperature intercooling system is advantageous for BSFC at all engine speeds. This requires small turbochargers to operate well at low gas mass flow.0 1. engaging the modified AC-system leads to a reduction in BSFC of up to 9%. 40 Brake torque in Nm 200 ∆T = 27. On the one hand side.0 0. Turbine in °C pcyl. This can be particularly advantageous for low speed operation as turbo charger surge is less likely to occur. rated power needs to be on par with the larger reference engine. On the other hand side. low end torque needs to be as high as possible to give the driver the impression of a larger displacement naturally aspirated engine. max in bar 80 Knock 900 70 850 60 e f 50 800 30 300 25 290 COHR in °CA be in g/kWh 20 280 15 270 10 260 g h 5 250 1500 2500 3500 4500 1500 2500 3500 4500 5500 Engine speed in min-1 Engine speed in min-1 Fig.5 K 170 with AC Compressor 10 a b 160 0 20. This asks for large 9 . the torque band could be spread out further and output could be increased. 10 . It is assumed that the torque increase can be achieved over the rpm- range of the nominal max torque. At full nominal engine load the system showed advantageous properties regarding BSFC at all engine speeds. 5: Results for improved output tests 4 SUMMARY OF RESULTS AND IMPROVEMENTS IN SYSTEM SETUP Overall.turbochargers. This is particularly the case in situations where the engine is operating with enriched air-fuel ratio for thermal component protection. Additionally. This could be achieved through a density increase in the intake air and improved thermal efficiency by earlier COHR. In some of those cases. The nominal max torque range could be extended from 3500rpm to 4500rpm. It could be demonstrated that the concept can aid in increasing low-end torque of a downsized turbocharged engine.5%.5 % 200 1600 180 160 1400 140 120 1200 30 950 900 TExh. it could even be increased by 12. As Fig. the availability of colder intake air also aids the dynamic behaviour of downsized engines. low temperature intercooling can ease the conflict between those demands. it was found that low temperature intercooling employing a modified vehicle AC system can improve efficiency by up to 9% at some high load points of engine operation. 5 shows. 240 1800 Boost pressure in mbar Brake torque in Nm 220 12. With an intake temperature reduction of about 30K. None of this needed modifications of the turbocharger. None of these results needed turbocharger modifications. If stoichiometric combustion is possible with regular intercooling. Also. No damage or abnormal wear was found in the engine after the test trials. Turbine in °C 25 COHR in °CA 850 20 800 with AC Compressor 15 w/o AC Compressor 750 700 10 650 5 600 1000 2000 3000 4000 5000 1000 2000 3000 4000 5000 Engine speed in min-1 Engine speed in min-1 Fig. As shown in [9]. the additional power needed to drive the low temperature cooling system can be larger than the energy saved by a more efficient COHR. At 1750rpm. the system shows no advantages. boost pressures required for nominal full load could be reduced. the brake torque at 1500rpm was increased by 11% without increasing boost ratio. the AC compressor’s efficiency is highly dependent on the vaporization temperature and pressure. It would also increase the efficiency of the cooling system significantly due to the elimination of losses of one heat exchanger step. 6: Improved system setup with intercooler / evaporator (ICev) [10] The cold fluid circuit employed in the original experimental setup should be eliminated. space and weight for additional parts and fluids. an expansion valve and some lines. Besides gains in efficiency. For a real world application. If the efficiency of the low temperature intercooling system (cooling power to mechanical power consumption) could be improved. the authors came to the conclusion that the basic experimental system already showed promising potential for successful introduction into niches of the automobile market. low-speed driving. On the other hand side. on the one hand side. Potential exists when COHR is later than ideal for efficiency. This would permit approximately 10°C charge air temperature at R134a vaporization temperatures of around 0°C to 2°C. charge air temperatures after intercooler can exceed 40°C. The system should be optimized as a whole towards its use for intercooling. A higher vaporization temperature means a better efficiency of the low-temperature cooling system. The experimental setup consisted of parts available “off the shelf” without any specialized development. This potential could be increased with refined integrated system design and development. In general. The whole parts effort for such a low-temperature intercooling system would consist of an evaporator-intercooler. Those temperatures are very similar to the vehicle’s AC system. Also. In those cases. due to the danger of knock or cylinder pressure limits. particularly in hot surroundings or during high-load. 6. This is. 11 . it is also expected that pressure losses could be reduced in the intake manifold. A production system should benefit from an integrated design of the AC components for passenger compartment AC and intercooling. benefits could already be gained at somewhat lower engine loads. the system investigated was a provisional experimental setup. the ICe should be an evaporator core without an in-between-liquid-system as shown in Fig.In real world driving. This would save cost. Fig. helpful to combine the systems and use synergies during parts development. the introduced concept can show even larger potential. REFERENCE LIST [1] KUBERCZYK.: Ladeluftkühlung durch Abgasenergienutzung .. SCHERER. BAAR. http://www. 2012 [4] WILLAND. 2012 [9] GUHR C. R.. F. SCHERER. ZEGENHAGEN.de/auto/2012-10/dudenhoeffer-motor-ps.Engine Operation at Charge Air Temperature below ambient temperature.: Verbesserung von Effizienz und Dynamik eines hubraumkleinen turboaufgeladenen 3-Zylinder-DI-Ottomotors durch Abgasrückführung und ein neues Ladeluftkühlkonzept. Dissertation TU Berlin 2014 12 .: Grenzen des Downsizing bei Ottomotoren durch Vorentflammungen.: Heat2Cool- Increased Gasoline Engine Efficiency due to Charge Air Cooling through an exhaust Heat Driven Cooling System. S. ZINNER.: Einfluss der Ladelufttemperatur auf den Motorbetrieb und ihr Potenzial auf die Steigerung der Leistungsdichte des aufgeladenen Ottomotors. MTZ 01/2014 (to be published) [6] KADUNIC. R. F. T. ZEGENHAGEN. ZIEGLER.. 2009 [2] GOLLOCH.: Downsizing Bei Verbrennungsmotoren: Ein Wirkungsvolles Konzept zur Senkung des Kraftstoffverbrauchs.Ihr Einfluss auf die Abgasemissionen..zeit... T.: Wir müssen das Rennen um die höchste PS-Zahl stoppen. und Dieselmotoren: Bewertung von wirkungsgradsteigernden Maßnahmen bei Ottomotoren. S. J. K. Springer-Verlag.: Aufladung von Verbrennungsmotoren.: Heat2Cool. S. MTZ 05/2009 Jahrgang 70 [5] KADUNIC. Berlin Heidelberg. F. 2005 [3] Dudenhöffer. Aachener Colloquium 2013 [7] SCHERER. BAAR. R. R. Springer Vieweg. Dissertation TU Berlin 2014 [8] PUCHER. H. F..: Wirkungsgradunterschiede zwischen Otto. Dissertation TU Dresden 2011 [10] KADUNIC. Expert-Verlag. F.. B A Gurunathan. UK R Jackson. A G J Lewis. UK ABSTRACT Current trend on engine downsizing forces engine manufacturers to contemplate powertrains with more than one boosting device. W S-I Wan-Salim. UK C Copeland. R F Martinez-Botas Imperial College London. Jaguar Land Rover Limited. bar] T Temperature [K] UB UltraBoost W Power [W] _______________________________________ © The author(s) and/or their employer(s).Assessment of supercharging boosting component for heavily downsized gasoline engines A Romagnoli. NOMENCLATURE Abbreviation Unit Subscripts Unit BSFC Brake Specific Fuel OUT Outlet Consumption INL Inlet BMEP Brake Mean Effective T Total Pressure S Static C Flow velocity [m/s] SC Supercharger cp Specific Heat TT Total-to-total Constant Pressure [J/kg·K] ER Expansion Ratio Greek HP High pressure ρ Density [kg/m3] LP Low pressure η Efficiency m Mass Flow Rate [kg/s] γ Specific Heat Ratio N Speed [rpm] PR Pressure Ratio P Pressure [Pa. L Matteucci Lotus Engineering. The presence of these devices leads to complex 1-D engine models which rely on performance maps provided by turbo/supercharger manufacturers. no detailed analysis has been carried out to understand how these maps affect engine performance simulation. C J Brace University of Bath. The acquired data were used to assess the effectiveness of 1-D engine performance prediction and to contemplate the opportunity to exploit the boosting system and use it as engine charge air cooler in the form of an expander. UK J W G Turner. As part of the UltraBoost project (65% gasoline engine downsizing). 2014 13 . N Luard Powertrain Research and Technology. So far. Imperial College tested the boosting components of a turbo-super configuration. S Akehurst. After an assessment phase. In order to meet the 400 200 targets outlined above. This supercharger is capable of running with high adiabatic efficiency over a wide range of operating conditions. The UltraBoost programme is an ambitious project aiming to deliver a 2.1 INTRODUCTION The need for more efficient. layout [2] At present. Engine models (1-D in nature) are being extensively used by engine makers since they provide simple. single speed drive. 1. The technical challenges associated with engine downsizing (typically of up to 40%) are well understood and OEMs are acting to develop engine concepts capable of real world fuel economy improvements whilst delivering sufficient vehicle driveability and performance. the analysis relied on the performance maps provided by the manufacturers [3]. The theory behind engine software is not new and the main advantage is that its development relies on extensive database which come from many years of Figure 2: UB engine system experimental activity and model optimization. a two-stage 100 50 series arrangement was selected.5 bar absolute and offering approximately up to 35% potential for the reduction of fuel consumption and CO2 emissions while still matching a JLR’s 5. these have been used as baseline curves to 500 250 ENGINE POWER [kW] build up all the analysis carried out in this paper.0 litre V8 naturally aspirated (NA) engine performance figures (65% engine downsizing). fast but still physically based tools for preliminary engine design. The target power and torque curves 600 300 ENGINE TORQUE [Nm] llloll are given in Fig. air pressure charging of up to 3. looking at many different 200 100 options (including electrification of some components). the output prediction from 14 . Data for the baseline engine platform where made available by Jaguar Land Rover whereas for the selection of the boosting components (turbocharger and supercharger). Torque Power This includes a LP stage (GT30R 0 0 turbocharger by Garrett) and a HP 0 1000 2000 3000 4000 5000 6000 7000 stage mechanical supercharger ENGINE SPEED (rpm) from EATON (Fig. a high level of boosting for the engine is 300 150 required. More specifically. The selection and matching of the boosting system was mainly supported by 1-D engine simulation (using GT-Power). for example through turbocharging [1].0 Litre four cylinder gasoline downsized demonstrator engine capable of up to 35 bar Brake Mean Effective Pressure (BMEP). more performing and less polluting engines is forcing OEMs to significantly downsize engine capacity. 2). the supercharger is a Figure 1: UB target Power and Torque Roots type Twin Vortices Series (EATON TVS R410) with a clutched. However turbocharger turbines present a long standing issue which is related with the width of turbine performance maps provided by turbocharger manufacturers. More in particular.1 Experimental set up All of the Ultraboost engine development work was undertaken in one of the University of Bath transient engine facilities. 5]. fuel and cooling circuits were all achieved through the CADET v-14 software [4. CAN instruments. conditioning of combustion air. surge. assessment of the impact of above-ambient inlet conditions on supercharger performance prediction in current 1-D engine software. forces the HP compressors to operate at different inlet pressures and temperatures to what they have been designed and tested for. Interfaces to the combustion analysis. the rapid electrification of powertrain also requires development of mathematical models for electric machines. 2 AIM AND OBJECTIVES The aim of this paper is to provide an insight into supercharging boosting technology in heavily downsized gasoline engines. In addition to this. HP stage declutched and driven by the pressurized air inlet instead of the engine crankshaft). EMS. turbochargers and superchargers represent two boundary conditions for the engine block since they determine the inlet pressure and temperature of the charge air feeding the engine.e. fuel flow. emissions analysers. Maintaining the AC dynamometer allowed motoring work to be undertaken for controller debugging and friction tests as well as improved transient response during time to torque testing. combustion parameters. the advent of sequential/series installation on the intake side. 3 EXPERIMENTAL ANALYSIS 3. specific experimental hardware was used for measuring engine speed. In addition to this. knock. In addition to standard temperatures and pressures. . All control and data acquisition was performed using a Sierra-CP CADET V-14 control system. two main aspects of supercharging will be investigated: . intake manifold and engine blowby. the HP compressor could be contemplated as an additional cooler for the engine charge air. Implementation: experimental analysis and comparison with GT-Power prediction.current models compares quite well with engine test bench data. Justification: when not in use. The facility features a twin dynamometer arrangement with a 220kW AVL AC dynamometer supplemented by a Froude eddy current brake to allow additional absorption up to the expected torque and power rating of the Ultraboost engine. Amongst the many areas which would require specific analysis. These maps (usually obtained with hot gas test stands) are narrow in range and force engine simulation software to rely on significant extrapolated map points. assessment of supercharger performance when run as an expander (i. motors and generators which are not always available in standard software packages. Again this poses a question on how well the HP compressors output parameters are predicted by current 1-D engine software. turbine maps extrapolation etc. the current paper intends to assess the impact on engine simulation output due to the presence of boosting devices. Justification: see Section 1. Implementation: experimental analysis. emissions. torque. In 1-D models. 15 . However the level of simplification intrinsically embedded in this software requires simulation engineers to calibrate these models in order to compensate for the lack of data and inadequate calculation routines in critical areas such as combustion. mass flow rate. a bespoke test facility was set up at Imperial College with a 100 kW electric motor (nominal speed of 1800 rpm) used as main drive. The pulleys/belt assembly on the supercharger side replicates the same layout as the UB engine using the same groove type pulley and belt (green box in Fig. only pressure and temperature were measured (Fig. a system of step-up gears was included in the rig design (Fig. At the exit to the supercharger. speed of 24000 rpm. 4). 4). Figure 3: Supercharger test rig overview As per the supercharger analysis. in order to achieve the required speed. At the inlet to the supercharger. 3 and 4). Figure 4: Supercharger testing layout 16 . static pressure tappings (using a Scanivalve system) and K-type thermocouples respectively. The speed of the motor is well below the nominal speed of the EATON supercharger which is rated with a max. Hence. 3). The power absorbed by the supercharger was measured with a torque meter positioned after the crankshaft pulley (Figs. pressure and temperature were measured using a V-cone (DP meter from ABLE). This is defined as the ratio between the consumed power of the supercharger (WSC. The full load curves at high engine rpm have not been completed due to some issues which arose at the time of testing (even though a full sweep for power and torque is currently being done and it will be material for future publications).The inlet to the supercharger was built in modular units allowing two types of tests. Eq. EGR (Exhaust Gas Recirculation) control strategy and water cooled exhaust manifold. At low engine rpm instead (from 1000 rpm to 1250 rpm). Option 1 set up was used to generate standard supercharger performance maps (refer to Section 3. The isentropic efficiency (total-to-total) was calculated using the thermodynamic correlations in Eq.2) whereas Option 2 was used to generate test data for 1-D engine simulation comparison (Section 4. This is due to the boost system not being able to deliver the target pressure as initially predicted by the simulation during the selection phase of the boosting components (for more details please refer to [3]). 5) provided by the electric motor (measured with the torque meter). Eq. Referring to the black box in Fig. The engine set up included the boosting components (HP and LP stage). Power 0 0 0 1000 2000 3000 4000 5000 6000 7000 ENGINE SPEED [rpm] ) Figure 5: UB Target and Experimental Power and Torque The HP stage was tested and its results compared with those provided by the manufacturer. the experimental torque shows significant deficit when compared to baseline engine (28% and 12% less torque and power respectively). 7. Besides the efficiency. The figure shows that the UB engine meets the requirements for power and torque. 6) and the input power (WSHAFT. 3. The UB engine was tested at full load conditions for a range of speeds going from 1000 rpm to 6000 rpm. 4. Since the pulleys/belt assembly after the torque meter replicates the same layout as that of the UB engine. an additional parameter (here defined as Power Ratio) was also calculated. Torque 100 Target Power 50 Exp.0L engine (Target Torque and Target Power).2 Experimental results In this section the test results for the UB engine and the HP stage (supercharger) are reported. Option 1 consists of letting the supercharger inlet exposed to ambient conditions (atmospheric pressure and temperature) whereas Option 2 allows the inlet pressure and temperature to the supercharger to be set by means of an inline heater and compressor. 600 300 500 250 ENGINE TORQUE [Nm] 400 200 ENGINE POWER [kW] 300 150 200 100 Target Torque Exp. above ambient inlet conditions) and to assess the supercharger performance when run as an expander (Section 5). In Fig. the Power Ratio is intended to represent a pseudo- 17 . 5 the measured power and torque curves for the UB engine (experimental Torque and experimental Power) have been compared with those obtained from the baseline 5. From Fig.02 0. / . 2. 6 it can be seen that the mass flow rate values agree well with those measured by EATON. The trend for the mass flow rate curves is consistent with the original maps from EATON and a difference in mass flow no larger than 4% could be found. = (8) The measured mass flow rate and efficiency results have been plotted in Figs.2 1 0.00 0. (6) ⁄ ( ) −1 . this could lead to larger clearances between the supercharger lobes and the casing (due to thermal expansion). 6 & 7 and compared with those provided by EATON.14 0. This can probably be attributed to the higher temperatures experienced within the supercharger at higher rotational speeds.4 1. − .4 2. + (3) = + 0.8 1.10 0.5 ∙ ∙ (4) 2∙ = ∙ (5) = ∙ . The Power Ratio values were then used to generate a look-up table for the UB engine model in order to determine the transmission losses occurring between the crankshaft and the supercharger.mechanical efficiency for the SP-EB-CB (green box in Fig. −1 . 3).12 0.6 8krpm 10krpm 12krpm 14krpm 16krpm 18krpm 20krpm 6krpm 2.06 0.04 0. = ℎ = (7) . Different considerations have to be made for the 18krpm and 20krpm where a larger discrepancy between the mass flow rates (black and red points) was found. = .6 2krpm 1.08 0.16 MASS FLOW RATE [kg/s] EATON IMPERIAL Figure 6: Validation for supercharger mass flow rate measurements 18 .2 PRESSURE RATIO [TT] 2 4krpm 1. This explanation was partly supported by the supercharger manufacturer (EATON) which pointed out the fact that the unit tested at Imperial College was a one-off prototype with potentially different coating material for the lobes and hence with different response to thermal variations (no further indication has been provided in this direction). This means that for supercharger speeds lower than 6000 rpm. The GT-Power analysis was done using a validated 1-D engine model for the UB engine (refer to Section 1). the engine would be running at speeds less than 1000 rpm. the Power Ratio values have also been reported (dashed lines). 100% 90% 14krpm 16krpm 18krpm 90% 12krpm 20krpm 85% 8krpm 80% 2krpm 80% 4krpm 6krpm 70% 75% 60% EFFICIENCY [TT] POWER RATIO 50% 70% 40% 65% 20krpm 30% 18krpm 60% 14krpm 20% 2krpm 12krpm 8krpm 55% 10% 4krpm 6krpm 10krpm 16krpm 0% 50% 0. For validation. operational speeds at which the supercharger will not be operating.2.As per the supercharger efficiency. 7).16 0.1 Supercharger performance analysis The supercharger was tested for rotational speeds.The speed ratio between engine crankshaft and supercharger speed is 5.06 0. However for the current testing programme.14 0. good agreement between the EATON and Imperial supercharger test data was also found (Fig.00 0. 4. In the secondary axis. The curves show that for low rotational speeds (2krpm to 6krpm) the power ratio values fall below 75%.18 MASS FLOW RATE [kg/s] EATON IMPERIAL Power Ratio Figure 7: Validation for supercharger efficiency measurements (primary axis) and plot of the Power ratio (secondary axis) 4 COMPUTATIONAL ANALYSIS The following paragraphs analyse the output of the GT-Power simulations for the HP stage performance. the results for the 1-D engine model was compared to engine test bed results as described in Sections 1 and 3. inlet mass flows.08 0.10 0. it was chosen to not exceed 2500 engine rpm on the test bed in order to avoid damaging the supercharger as a consequence of the large 19 .12 0.02 0. The control strategy for the UB engine. This aim is to assess how well GT-Power simulation calculates the performance parameters when the input values are different than ambient.88. with thermal efficiencies beyond 70%.04 0. pressures and temperatures equivalent to those obtained from the UB engine at full load conditions. forces the supercharger to operate up to 3500 engine rpm ( 20580 supercharger rpm). However for rotational speeds which are more representative of the supercharger speeds in real engine operating conditions the power ratio climbs up to 85%. SC efficiency). [Pa] 1.254 Inlet Pres.21 bar).8 10 0. pressure.inlet pressure and temperature.429/1. there is no clear pattern between the measured and predicted values.180/0.832/2. the test results show that there are some discrepancies between the measured and predicted data.0 360 1.09 0.0 370 2. From the table it is apparent that the speed. temperature and mass flow rate have been set accordingly. The comparison for the supercharger inlet conditions between the 1-D simulation and the tests is given in Table 1.1 Tests Imperial 0 0. Table 1: Supercharger inlet conditions (1-D simulation/Imperial Tests) NEngine [rpm] 1250 1500 2000 2500 NSC [rpm] 7350/7359 8820/8844 11760/11750 14700/14742 Mass flow [kg/s] 0. with a discrepancy no larger than 1%. the measured and predicted outlet pressures diverge significantly with a variation of about 6% (∆P=0. SC power.6 SC POWER [kW] EFFICIENCY 0.78 Inlet Temp.0 350 500 1000 1500 2000 2500 3000 500 1000 1500 2000 2500 3000 ENGINE SPEED [rpm] ENGINE SPEED [rpm] 12 0.4 4 0.5 365 2.057 0.73 2.2 2 1-D Simulation 1-D Simulation Tests Imperial 0.3 0.321 2. SC outlet temperature.172 0.5 6 0. As per the temperature.5 Tests Imperial SC OUTLET PRESSURE [bar] 375 3. 8 the measured outlet conditions to the supercharger have been compared with those obtained with the simulation (SC outlet pressure. The same occurs for the SC power which shows larger variations as the engine speed increases (at 2500 rpm. At higher engine rpm.0 385 1-D Simulation 380 SC OUTLET TEMPERATURE [K] 3. difference of about 20 .320/2.0939/0.7 8 0.5 1-D Simulation 355 Tests Imperial 1.734/1.9 0.265/0.060/0. [K] 303/302 320/320 333/332 347/353 4.434 1. Even though the trend of all the measurements is well captured by the simulation.0 500 1000 1500 2000 2500 3000 500 1000 1500 2000 2500 3000 ENGINE SPEED [rpm] ENGINE SPEED [rpm] Figure 8: Comparison between measured and predicted SC data In Fig. If there is energy available to over-boost an air charge.53 kW). if the heat of compression is removed via a standard intercooler prior to expansion. intake air pressure is raised by a compressor in order to increase air density and therefore. Perhaps most notably. While there are other ways to improve knock tolerance in a pressure charged SI engine (increasing EGR rates and improved port and combustion chamber design). this additional pressure can be expanded over a turbine (expander) to remove heat. 7) when inlet temperature is higher than ambient is no longer appropriated. A variation in outlet temperature of ±5ᵒC can lead to an efficiency variation of several percentage points. 5 SUPERCHARGER AS AN EXPANDER 5.30. ∆Power=3. since heat transfer through the casing affects the SC exit temperature (which could partly explain the disagreement observed between the predicted and measured outlet temperatures). Due to the potential advantages of this approach. The concept and the T-S diagram from this work are shown in Fig. Heat transfer effects on compression process for centrifugal machines were explained by the lead authors in [6]. This increased air temperature can lower the knock limit of the engine .1 Turbo-expansion: initial simulation study In pressure-charged engines. Calculating the supercharger isentropic efficiency (using Eq. The idea is a simple one. The concept of turbo-expansion is not new – indeed it has been used widely to provide air conditioning to commercial aircraft for many years. the results are believed to be strongly biased by heat transfer effects.7%. Therefore. thus making intercooling necessary to reject this heat to the environment. As per the efficiency. the NOMAD engine development project by Lotus Engineering [7. since heat exchangers have effectiveness below unity. lowering charge temperature at start of compression could provide a direct route to improved fuel economy. 8] sought to demonstrate the advantages of turbo-expansion using an Opcon mechanical twin-screw expander. However. the charge air temperature is also raised by the compression process. it is possible to cool the air charge below ambient temperature. 9. Both of these mechanisms for knock control can lead to compromises on fuel economy for a given BMEP target. specific power output. Figure 9: Turbo expansion process [7] 21 . the temperature of the intake air will always be above ambient temperature in a pressure charged engine. However. there have been recent attempts to apply this concept to the air path of a SI engine.making it necessary for lower compression ratios or greater spark retard. there are operating points where there is excess turbine energy that could be used to generate additional boost by modulating the wastegate.Whelan et. the full-load BSFC was predicted to improve by 4-5% due the sub-ambient air charge from the turbo-expander. the inlet to the supercharger was pressurized with a compressor and allowed to expand from inlet to exit. Figure 10 shows the expansion ratio versus mass flow for the Eaton device operating as an expander at constant speeds of 2000rpm. al. there is no available performance test data of an Eaton supercharger acting as an expander. Although it was not explored in this study. As an example. The turbo-expander compressor provided final pressure ratio of 1. While this dual purpose is not possible with a fixed ratio drive.25) compared to standard operation. 5. diminishing returns could be expected for reduced expansion efficiencies as noted by the work by Turner et. the possible switching between compressor and expander operations could be envisaged using such a variable ratio drive system. could the Eaton supercharger be used as an effective expander? While this would require additional boost from the turbocharger. 4. This approach was investigated using the UltraBoost. Since. Despite the additional parasitic loss of the supercharger (MEP= +0. In the model arrangement the additional components were placed within the two stage supercharger/turbocharger air path.al. in order to deliver a similar engine mass flow. [8]. It can be immediately noted that to deliver a similar mass flow range to the engine. This brief study considered a single engine speed at 2000rpm where the supercharger supplied an over-boost (pressure ratio +0. The temperatures and pressures were recorded as before and the resultant torque and speed was measured. the experimental facility at Imperial College was used to test this capability. the expansion stage will lead to sub-ambient temperatures. to the authors’ knowledge. the supercharger would have to spin at approximately half the speed compared to the operation as a compressor.25 before rejecting heat through a heat exchanger (ε=0. model-based research into the use of a continuously variable transmission (CVT) supercharger has been carried out by the University of Bath in reference [11]. GT-Power model as part of an initial study into the possible benefits of turbo- expansion to a highly downsized SI engine. For reference. 4000rpm and 8000rpm. By removing heat between the compressor and turbine stages with a second charge air cooler. Thus. 10] proposed a slightly different turbo-expansion concept where a compressor and turbine are both placed in the intake flow path to raise and lower the intake pressure respectively. In fact. this posed an interesting question.000rpm. In addition.2 Eaton Supercharger as expander The GT-Power simulation suggested that there could be significant benefits from turbo-expansion air charge cooling. this was more than offset by the benefit of being able to advance ignition timing due to the lower charge temperature at the end of the compression stroke. The turbo-expander performance was provided from scaled turbocharger maps which generated total to static efficiencies of ~75%. 22 .4 is desired.8bar). a mechanical connected expander could present an indirect means to recover exhaust gas energy. As shown in Fig. namely. since the UltraBoost arrangement places the supercharger downstream of the turbocharger. the supercharger must spin at a slower rotational speed if it is to operate as an expander. the supercharging characteristics are also presented as delivering a ‘pressure ratio’ for speeds up to 10. [9. if an expansion ratio of 1.5).85) and expanding through the fresh-air turbine (expansion ratio ~1. In this situation. 23 .2 1.11 0.8 EXPANSION RATIO PRESSURE RATIO 1.0 1.09 0.0 0.05 0. if an expansion ratio of 1.4 1.6 1.08 0. 9.04 0.0 2.4 is desired. 12 demonstrates the power that can be generated. / .01 0. Fig. operating as an expander. and thus returned to the engine. . if it is operated as an expander.00 0. Finally.07 0. the supercharger can only deliver isentropic efficiencies between 40-50%. The definition of expansion efficiency is simply taken from a turbine definition as shown in Eq.2 2. The values are also reasonably consistent with those reported in reference [8]. Picking up the example earlier.10 0. turbo-super arrangement if it were possible to change the drive ratio.4 1.0 1.2 1.12 0. These results represent an intriguing new use of an Eaton supercharger and show that the device could be used as an expander in the series. It is this rather poor expansion efficiency that may lead to diminishing returns from a sub-ambient intake air charge.6 1. further modelling work must be carried out using the data presented here. approximately 1kW of power can be generated while simultaneously reducing the air charge temperature. = ⁄ ℎ = (9) 1 − (1⁄( ) ) . 2. it is also clear that.2 2. 1− .03 0.02 0.13 MASS FLOW RATE [kg/s] Imperial_4000rpm Imperial_6000rpm Imperial_8000rpm Imperial_10000rpm Imperial_12000rpm SC_Expander_6000rpm SC_Expander_4000rpm SC_Expander_8000rpm Figure 10: Pressure and Expansion Ratio of Eaton Supercharger versus Mass Flow Figure 11 demonstrates the total-to-total efficiency of the supercharger as an expander compared to the compressor operation. To understand the trade-off between additional turbocharger boost (turbine backpressure) and the benefits from turbo-expansion across a supercharger.06 0. This is not surprising considering the design intent of the device was not to operate in this way. However. It is immediately clear from this data that a substantial drop in isentropic efficiency is to be expected from an attempt to use the EATON as an expander.8 1. The analysis focussed on the HP stage and is aimed at understanding the effectiveness of commercial 1-D engine software when predicting the performance of boosting elements (a supercharger in this case) with inlet conditions other than ambient (pressure and temperature beyond atmospheric).200 1.80 EFFICIENCY [TT] SC_Expander_4000rpm SC_Expander_6000rpm SC_Expander_8000rpm Imperial_4000rpm Imperial_6000rpm Imperial_8000rpm Figure 11: Total to total isentropic efficiency of the supercharger operating as a compressor and expander 1.000 0.800 1.700 1.700 1.600 1.200 1.500 1.300 1.000 1. The engine boosting system layout is a two-stage series layout.500 1.600 EXPANSION RATIO PRESSURE RATIO 1.100 1.50 0.000 -2000 -1500 -1000 -500 0 POWER SHAFT[W] SC_Expander_2000rpm SC_Expander_4000rpm SC_Expander_6000rpm SC_Expander_8000rpm Figure 12: Power harvested from the supercharger operating as an expander 6 CONCLUSIONS The current article describes the outcomes of a research conducted on a heavily downsized gasoline engine (UltraBoost project.70 0.20 0.800 1.10 0.100 1. with a LP stage (turbocharger) and a HP stage (EATON TVS supercharger).300 1.400 1. 65% engine downsizing). 1.800 1. 24 .300 1.500 1.400 1.400 1.00 0.700 1.600 EXPANSION RATIO 1.30 0.60 0.40 0.200 1.100 1. Shell Global Solutions. discrepancy of 6% encountered at 2500 rpm only). UK. S. G. Copeland. P. Pearson. China. C. R. S. Luard.uk – Powertrain Vehicle & Research Centre (PVRC). the difference between measured and predicted values can be significant for some engine rpm. 10th IMechE International Conference on turbochargers and turbocharging. [3] Copeland. Coventry. [4] Karl. J. Richardson. M.. Sam. L.... CD-Adapco. A. F2012-A01-021. Carey. Boost System Selection for Heavily Downsized Spark Ignition Prototype Engine. nick. 2012. University of Bath and University of Leeds) for the support provided. C.e. Martinez-Botas. Andrew. Carey. an experimental study on the supercharger performance was also carried out in order to assess the opportunity to use the supercharger as an additional charger air cooler (i. Robinson. R. R. An agreement within 1% could be achieved for the inlet conditions between simulation and experiments. GE Precision Engineering. 2012. the test results also showed that a non-negligible amount of power (more than 1kW) can be generated from the supercharger when run as an expander. C. As per the supercharger efficiency. the simulation and experiments show similar trend even though. supercharger was run as an expander). The experiments showed that there is a substantial drop in the supercharger isentropic efficiency when used in this way..bath... London. [2] UB200 Boosting Layout. [5] Online source. Martinez-Botas. A. As per the supercharger output values. 25 . Beijing. Chris. Di Martino. Turner. University of Bath website: http://www. the test results showed that the 1-D engine software provides a good prediction for the supercharger outlet pressure (max.7% higher than that predicted at 2500 rpm. These results are quite interesting and could potentially open the way to consider a supercharger as an enabler for enhanced engine fuel economy. L. In addition to the supercharger simulation prediction analysis. Salamon. C. P.ac. Richardson. no valid comparison could be brought forward since the measured efficiency was found to be significantly biased by heat transfer due to the heated air inlet temperature. besides serving as a charge air cooler. FISITA 2012 World Automotive Congress. C.. Chobola. J. R. Romagnoli. the supercharger could also be contemplated as a power generation system. 21st Aachen Colloquium Automobile and Engine Technology. Jaguar Land Rover.. inlet pressure and temperature as those predicted by the engine model at the inlet to the supercharger.. On the other hand. REFERENCE LIST [1] Improving Fuel Economy by 35% through combined Turbo and Supercharging on a Spark Ignition Engine. Turner.The EATON TVS supercharger was tested with similar speed. JLR Technical Presentation.. N. A penalty of more than 20 percentage points could be found (efficiencies less than 50%) for the speeds under exam (supercharger speeds 4000 rpm to 8000 rpm).. Mc Allister. As per the outlet temperature and supercharger power. For instance the measured supercharger power was found to be 30. This suggests that. June 2011. ACKNOWLEDGEMENTS The authors would like to thank the UK Technology Strategy Board (TSB) for funding the UltraBoost project and all the consortium partners (Jaguar Land Rover.. 2012. The effect of advanced combustion control features on the performance of a highly downsized gasoline engine.. The operating conditions considered for the supercharger analysis were equivalent to the engine running at full load conditions with speeds varying from 1000 rpm to 2500 rpm. B. Lotus Engineering. M. Sept. 2012. S. R.... Eng.. and Oscarsson. C.1177/0954407013492932. Pages:58-77. 2005..... Pearson... Spence. D. SAE paper no.. Richards. [10] Whelans. C. “The Turboexpansion Concept .[6] Romagnoli A. Bassett. R. [8] Turner. “Turbo-cooling applied to light duty vehicle engines”. Richards. [7] Turner.. and Brace. DOI: 10.. R. and Blundell. Young. “Investigation into the trade-off between the part-load fuel efficiency and the transient response for a highly boosted downsized gasoline engine with a supercharger driven through a continuously variable transmission” Proc.. A. R. Vol:38. 2003-01-0401. Pearson. A.. Heat transfer analysis in a turbocharger turbine: An experimental and computational evaluation. 2013.. [11] Rose. W. Martinez-Botas RF. 2005-01-1853.. Bassett. Akehurst S. J. 1359-4311. 2010.Initial Dynamometer Results”. “Design and Development of a Turbo-expander for Charge Air Cooling” IMECHE Turbochargers and Turbocharging. SAE paper no.. J. “Performance and Fuel Economy Enhancement of Pressure Charged SI Engines through Turboexpansion – An Initial Study”. C. 26 . Part D: Journal of Auto. Applied Thermal Engineering. of IMECHE. J. Congress. [9] Whelans. M. By combining the characteristics of the downsized turbocharged engine and the scavenging process of a more-normal naturally aspirated engine. some novel techniques need to be investigated. both the torque performance and the gas exchange process could be enhanced. In order to simultaneously enhance the already competitive advantages while mitigating inherent deficiencies of turbocharged engines. C Brace1. S Akehurst1. the DEP concept will be investigated in simulation using a validated highly downsized 2. J W G Turner2 1 University of Bath. UK 2 Jaguar Land Rover Limited. negative PMEP and knock sensitivity at high engine speeds are some of the main challenges facing downsized turbocharged SI engines. Divided Exhaust Period (DEP) is such a gas exchange concept where two exhaust ports from each cylinder are separated into different manifolds. BSFC and the stability of the engine were all improved due to the fact that the DEP concept features better gas exchange process and improved combustion efficiency. In this paper. Traditional downsized engine can currently only address the above problems within a very small window inevitably leaving the other situation compromised which would either affect fuel efficiency or engine power. while the other valve path (termed scavenge valve) bypasses the turbine to scavenge the remainder of the exhaust. The final results showed that the BMEP & transient performance. 2014 27 . C Copeland1. ABBREVIATIONS DEP: Divided Exhaust Period BSFC: Brake Specific Fuel Consumption SI: Spark Ignition PMEP: Pumping Mean Effective Pressure BMEP: Brake Mean Effective Pressure RGF: Residual Gas Fraction GA: Genetic Algorithm ___________________________________________ © The author(s) and/or their employer(s).0 Litre SI engine model. UK ABSTRACT Cylinder scavenging. while torque deficiency is more of a problem related to engines under low speed conditions. The blow-down pulse is directed through one valve that leads to the turbocharger in order to boost the intake charge.The effect of divided exhaust period for improved performance in a highly downsized turbocharged gasoline engine B Hu1. it is also attributed to the improved PMEP [3][4][5][6][7][8]. it is anticipated that the gas exchange process could be improved while keeping or exceeding the required BMEP. Improved combustion phasing from reduced RGF also results in reduced exhaust gas temperature and thus may reduce the need for other exhaust gas temperature control actions such as over-fuelling. Meanwhile. less RGF would cause the engine less prone to knock which then would lead to the advance of the spark timing [2]. This is due to the fact that DEP concept engine generally features a smaller turbocharger and for most of operating points. 2 TECHNICAL CONCEPTS The Divided Exhaust Period (DEP) concept is a novel way of accomplishing the gas exchange process in turbocharged engines with the aim to combine the positive effects of using a turbocharger while removing some of the negative aspects by using a Variable Valve Timing (VVT) system on the exhaust system. by bypassing the turbine a positive pressure gradient across the intake and exhaust system could be achieved. It should be noted that better gas exchange process and improved combustion phasing are the two major reasons to keep knock intensity under appropriate constraints for higher BMEP target. The blow-down valve functions just like the standard exhaust valve discharging the first portion of exhaust gas into the turbine to boost the intake air. However. It is known that RGF is strongly related to the knock intensity and engine test variability. the DEP concept could also benefit from improved combustion efficiency. By combining the characteristics of standard turbocharged and naturally aspirated engines. there are several obstacles such as cylinder residual gas scavenging. Specific reasons from the perspective of better BSFC and enhanced BMEP are as follows: BSFC benefit As the high backpressure between the turbine inlet and the exhaust port is isolated from the engine piston during the latter portion of the exhaust stroke. For fixed BMEP condition. Some author also stated that at knock-limited points. Hence. the adoption of water cooled exhaust manifolds or increased exhaust gas recirculation (EGR). Figure 1 shows the schematic view of the DEP concept. Conventional approaches to addressing these issues are to improve the engine side or turbocharger side efficiency. This is beneficial for the cylinder scavenging process resulting in a large decrease for Residual Gas Fraction (RGF). two valves are separated to feed the turbine and the exhaust pipe respectively. While the scavenge valve which evacuates the latter exhaust portion directly into the exhaust pipe (lower backpressure) aims at better gas exchange. Unlike the standard layout of the engine exhaust system. negative PMEP due to turbine induced back pressure. knock sensitivity and poor transient performance that are challenges facing this technology [1].1 INTRODUCTION A general trend to downsize engines means more and more SI engines feature turbochargers to improve fuel efficiency and BMEP performance. an increased PMEP could be attained. increased PMEP indicates less Gross IMEP thus better BSFC. 28 . This paper however focuses on a novel means of gas exchange process utilising Divided Exhaust Period (DEP) to separate the exhaust process into two distinct stages of blow-down and scavenge. BMEP benefit Enhanced BMEP could be achieved for the DEP concept engine if power density is a more important concern. Figure 1: Schematic view of the exhaust system layout in DEP concept 3 METHODOLOGY AND CONTROL STRATEGY In this paper.0 450.00 6.00 2. In order to understand the underlying gas exchange process. the timing of the scavenge valve (one degree of freedom mechanism) effect on BSFC and BMEP were investigated first.00 Valve Lift (mm) 8. For one degree of freedom study.0 360. the timing of the scavenge valve is also the control parameter to maintain or achieve the target BMEP.0 540.0 630. which is more like a traditional calibration.00 90.00 10. the function similar to a traditional wastegate. Optimization of the fully flexible mechanism was then undertaken and presented later in the paper (only for BSFC optimization situation). The reason for selecting the timing of scavenge valve to be the control parameter is due to the following reasons: 29 . the profile of the intake and the exhaust valves were retained to investigate the BSFC or BMEP improvement potential with minimum change of the whole system.0 180.00 A B 0. 12.0 270.00 4. The value of the other parameters will be based on the optimization for different purposes and will form a map in the ECU. seven degrees of freedom in the actuation of the DEP concept engine including the timing of the intake. blow-down and scavenge valves and the lift & duration of the blow-down and scavenge valves were investigated.0 Crank Angle [aTDC] Blowdown Scavenging Intake Figure 2: Exhaust valve timing It should be noted that in this paper. 30 . To be more specific. Variable valve timing (VVT) mechanism is more easily manufactured than variable valve lift and variable valve duration mechanisms. Meanwhile. less pumping work is needed thus fuel efficiency is increased. then the scavenge valve generally should be retarded. for some operating points. each degree of freedom represents a gene in the chromosome and BSFC forms the fitness function. *This strategy only applies to the high and some of the low to medium conditions. The matching of the IC engine and the turbocharger was performed by scaling the mass flow axis of the standard engine’s turbocharger maps. the benefit from the combustion stability is small. More detail about the structure of the GA system can be referred from figure 3. This strategy can also benefit from less blow-through which leads to less fuel injected. Some mathematic approaches for optimizing multi-variables systems and multi-objective responses thus need to be investigated. Fully flexible mechanisms of the intake and exhaust valves are therefore necessary to be introduced to exploit the maximum potential of the DEP concept engine. the profile of the scavenge valve needs to be shifted right to allow more overlap as shown by B in figure 2. It should be noted that potential PMEP gains due to the implementation of the DEP concept could offset some of the turbocharger power loss due to the smaller amount of exhaust mass flow into the turbine. changing one would cause the others to change their action on the engine system. If achievable BMEP is preferred. Multiple degrees of freedom mechanism Due to the possible large blow through and the limitation of only one controllable degree of freedom. the overlap between the scavenge valve and the intake valve accordingly alters which influences the amount of blow-through or short circuit air flow. a smaller sized turbocharger was selected by using mass multiplier in GT-Power [9]. the results might not be optimized as can be seen in the following section. However. As all the degrees of freedom are strongly interrelated. a smaller sized turbocharger may need to be selected to optimize the whole system. However. The operation strategy depends on the requirements of the system*. Strategy at the low speed is detailed in the RESULTS. because of RGF doesn’t improve significantly as in the situation when the timing of the scavenge valve retards. By advancing the scavenge valve. the relationship between the blow- down and the scavenge valve varies to feed a specific amount of the exhaust mass flow into the turbine. Thus. The scavenge valve is subjected to less thermal and pressure shock than the blow-down valve One degree of freedom mechanism By adjusting the timing of the scavenge valve. Without considerable effort. Overlap A and overlap B as can be seen in Figure 2 is necessary to avoid choked flow and to control the blow-through/or back flow respectively. A genetic algorithm has the capability to mimic the process of natural evolution and is thus considered to be one of the most useful approaches to solve this kind of problem. but the corresponding anchor angle (CA50) retards and the larger amount of fuel mass flow rate would adversely affect the fuel economy. the initial BMEP or desired higher BMEP is still not attainable. The schematic diagram for optimizing the results is shown below. If fuel efficiency is pursued. the scavenge valve should be set to a maximum advance that attains BMEP target. To be more specific. This would accordingly cause more mass flow into the turbine which increases BMEP and less RGF due to the larger amount of blow through. For strategy 2. The diameter of the valves was increased by 4mm to avoid choking when only one valve is involved in the gas discharge process. Strategy 1 for BSFC purpose In this strategy. The Douand & Eyzat knock model was implemented in GT-Power in an effort to predict knock onset for individual cycles [10]. firstly we aimed at minimizing the fuel consumption while matching the BMEP performance of the base engine.7bar were presented. It should be noted that the comparable results are from the original engine model in WOT condition. the aim is to reduce fuel consumption.9bar. The wastegate control strategy was replaced by the timing of the scavenge valve to maintain or exceed BMEP for strategy 1 and 2 below. the potential of exceeding the original engine performance was investigated. For strategy 1. Steady state operating points of 6500RPM/26. 4500RPM/30. residual gas content and to improve boost control. Figure 3: Genetic Algorithm Procedure 4 RESULTS This study was carried out on a validated 2 Litre extremely downsized SI engine model using GT-Power.2bar. It is worth mentioning that the target BMEP of this engine can be in excess of 32bar and it is anticipated that the gas exchange process is potentially significantly different to that of the engines other literatures investigated given their lower BMEPs. 31 . one degree of freedom effects were presented first with detailed explanation. In this paper. The reason why these four sets were chosen is because they represented four different gas exchange process detailed in the following. After that. Then multiple degrees of freedom interactions were presented with some explanations of the process occurring.9bar and 1500RPM/12. only one degree of freedom effect on BMEP was investigated. 2500RPM/17. Table 2: Averaged results for 4500RPM 4500RPM/30. This correlated with RGF very well with RGF decreased from 3.3 -2. DEP concept engine only needed39.3 1. It should be noted here.94. Unlike 6500RPM situation.0 which contributed more than the improvement of PMEP although it was not verified quantitively.72% while keeping BMEP fixed.3 to 19.7%.0 0.0 DEP 30.5 -1. As a wastegate was not required for boost control.2 258.75 17.49 23. This was largely due to the improvement of the pumping work as negative PMEP was increased by 44. the supercharger was removed.9kW.2 254. It is also anticipated that the DEP concept engine would also benefit from more stability due to the significant RGF decrease. Invalidated knock model could also influence the relationship between knock sensitivity and combustion phasing.9 246.2bar: BMEP (bar) BSFC (g/kW*h) RGF CA50 PMEP (bar) Turbine Power Original 26. This would further improve engine efficiency [3]. as can be seen in table 2.9 At 4500RPM. both the reduction of the pumping work and the improvement for combustion efficiency contributed the decrease of BSFC. The spark-timing advanced by about 4 degrees from 23. which indicated that more power was wasted to overcome the pumping resistance for the baseline engine.9. BSFC dropped a lot from 246.450 35.7 0. BSFC at 2500RPM and 1500RPM did not change too much which indicated that the DEP concept did not have significant BSFC benefit within lower engine speed.49 to 0.074 29.9bar: BMEP (bar) BSFC (g/kW*h) RGF Anchor Angle PMEP (bar) Turbine Power Original 30.1) One degree of freedom mechanism Table 1: Averaged results for 6500RPM 6500RPM/26.7 to 242.94 19. the original model is equipped with a supercharger as well but in order to investigate the DEP concept on turbocharged SI engines. it can be seen that the BSFC may be reduced by 1.9 DEP 26. At 1500RPM.2 39.3 -0. Lower pumping work can also be implied from the reduced turbine power. Possible reason could be poorer choking condition when only one exhaust valve is involved which adversely affects the temperature and pressure in cylinders. 32 .2 From table 1.7 2.0 3.2kW power to boost the intake pressure in order to attain the required BMEP while the base engine required as large as 43. the saved pumping loss during the wastegate open also contributed the improved PMEP. The combustion phasing on the other hand didn’t change too much which did not correlate with the reduction of RGF [9]. the increased BSFC was attributed to the decrease of the pumping work.9 249.2 43.74 17. 1 239. However for DEP concept engine.7 -0.7bar: BMEP (bar) BSFC (g/kW*h) RGF Anchor Angle PMEP (bar) Turbine Power Original 12. 33 . Figure 4: Pumping loop in P-V diagram at 4500RPM It can be seen from figure 4 that for the end of the exhaust stroke.2 5. the original engine continued to push the remaining exhaust gas from the cylinder to the turbine which increased the backpressure further and as the exhaust valve started to close there was a large peak in the cylinder pressure.105 5. Table 3: Averaged results for 2500RPM and 1500RPM 2500RPM/17.036 5.006 0.2 1500RPM/12.4 4.14 11. Pumping work can also be implied from the enveloped area.045 0. the cylinder pressure dropped dramatically to nearly atmospheric level which removed the backpressure to the piston.23 13.7 247. it is clear that the DEP concept benefit from less pumping loss than the original model.7 246.4 0.2 239.9 DEP 12.3 -0.4 3.36 14.3 4.6 DEP 18.31 10. as the blow-down valve was closed while the scavenge valve started to open.7 Pumping Loop Fuel consumption reduction between the standard and the DEP model can also be presented using the pumping loop in P-V diagram below.9 0.9bar: BMEP (bar) BSFC (g/kW*h) RGF Anchor Angle PMEP (bar) Turbine Power Original 18. 26 -0.14 4. it is almost impossible to achieve such improvement. The reason why the BSFC trend in fixed BMEP was not the same is attributed to the fact that at high speed with increased scavenge valve timing.25 PMEP (bar) -2.9 30. Strategy 2 for BMEP purpose This strategy targets higher BMEP.2 -1. Only the results of 4500RPM and 1500RPM were presented for simplicity purpose. the BSFC performance for all speed range was improved with large decrease of BSFC occurred around the high speed. However.4 12.9 0.7 252.3 19. due to the complexity of utilizing seven degrees of freedom mechanism for series production engine.045 0.2 245. Due to the improved gas exchange process and better combustion phasing.2 26.4 11. However.3 239. The decreased trend of BSFC for lower speed was basically due to the gain of PMEP. But it should be noted that even though BMEP trend is similar for these two sets of speeds and loads the curve of BSFC & RGF & PMEP is totally different.3 16.93 Anchor Angle 17.156 0.1 18. which can be seen from the last plot of figure 5.72 5.85 Table 4 shows the averaged results for multiple degrees of freedom mechanism.9 249.0 5.2 -0.9 38.062 Turbine Power (kW) 43. Figure 5 shows the relationship between the scavenge valve timing and BMEP & BSFC & RGF & PMEP.3 14.0 0.From the results above.1 35. The upper two plots indicated maximum BMEP potential at higher engine speed whilst the lower representing the lower speed BMEP trend.450 0. there was larger blow through across the cylinder. 2) Multiple degrees of freedom mechanism Table 4: Averaged results for multiple degrees of freedom mechanism 6500RPM 4500RPM 2500RPM 1500RPM Original DEP Original DEP Original DEP Original DEP BMEP (bar) 26. The performance of both situations improved by around 10% maximally. 34 .105 0. thus the behaviour or the performance of the engine should also be in the between.2 23.0 239.74 1.7 BSFC (g/kW*h) 258. thus the BMEP curve is more important than the corresponding BSFC.6 246. at low speed as the overlap between the scavenge valve and the intake valve was relatively small.49 1.0 29. Engines equipped with DEP concept in the future is most likely to be in the between of the one degree freedom and the seven degree freedoms.6 5.9 18.36 2.2 30.1 12.6 RGF 2. it can be summarized that for fixed BMEP condition.50 4.7 11. BSFC improvement depends on the trade-off between the benefits of the reduction in pumping work and the combustion efficiency gain.7 12.19 3. blow through effect was not noticeable.4 235. It can therefore be concluded RGF for DEP concept model not only depends on the blow through but also on the possible back flow. the blow through effect causes RGF decrease dramatically which is the case for higher speeds. Figure 5: BMEP. For comparative purposes. only the scavenge valve timing of 270 and 310 degree representing the lowest and largest RGF were selected. RGF and PMEP for 4500RPM and 1500RPM RGF curve difference can be explained by the following points: Timing relationship difference between the intakes. BSFC. However when the intake pressure is smaller as at the lower speed. the overlap between the scavenge valve and the intake valve is increased. In order to explore the gas exchange process in lower speed range. 35 . Figure 6 indicates the backflow for larger scavenge and intake valve overlap (lower blow-down and scavenge overlap). back flow effect would dominate [11]. For higher speed the first effect contributes more while for lower speed the latter one dominates. This can explain the RGF trend for lower engine speed well. blow-down and scavenge valve Pressure difference across the cylinder As with increased scavenge timing. Two different gas exchange process could occur here. the following figure was drawn. When the intake pressure is larger than the exhaust pressure when the scavenge valve opens. Large Back Flow Figure 6: Gas exchange process for different scavenge valve timing at 1500RPM Strategy 3 for transient performance purpose Transient performance is one of the major challenges to address in order to see widespread adoption of turbocharging in most production engines. Although further research work in this area is needed. 36 . This is majorly due to the fact that smaller inertia of the modified turbocharger is adopted. it is expected that DEP concept could benefit for transient performance of turbocharged engine. In addition to that. blow-down and scavenge exhaust temperature From figure 8. This is due to the fact that during some of the exhaust stroke and intake stroke. Figures 7 shows the comparison of the original. lift and duration of the exhaust valves. Higher blow-down but lower scavenge pulse temperature should be anticipated. the knock model for running the simulation which decides the anchor angle (CA50) might not be suitable for the DEP concept. This would either increase the cost of the system or affect the fuel efficiency. 37 . The choking condition could be largely improved by optimization of the timing. 1500 1000 Temperature (K) Original Blow-down 500 Scavenge 0 6500 4500 2500 1500 Engine Speed (RPM) Figure 7: The comparison of the original. It should be noted that for this study. The potential of this novel theory could be better proved with some testing results. only one of the exhaust valves was involved. This can also partly explain why PMEP drops for higher speed range when the timing of the scavenge valve retards. choking condition still existed and even poorer than the original engine. which would be verified in the future. the modified model for the DEP concept does not calibrate. it can be seen that even though the diameter of the exhaust valves were increased. Higher blow-down exhaust temperature and choking are the two major drawbacks for DEP concept engines. As the majority of high enthalpy mass flow is evacuated during the first portion of the exhaust stroke and a relatively low enthalpy mass flow which directly links to the exhaust doesn’t balance some of the blow-down pulse temperature.5 DISCUSSIONS It should be noted that even though the base engine model is validated using prototype engine testing data. blow-down and scavenge exhaust temperature when one degree of freedom involved in fixed BMEP condition. High temperature-resistant material should be used or fuel-enrichment strategy needs to be involved. the limit of the inlet turbine temperature was not considered. However. 7 ACKNOWLEDGEMENTS The authors would like to thank Nick Luard in Jaguar Land Rover Limited for his kind support in the GT-Power modelling. Blow through is dominating for high speed operating point while backflow effect is large enough for low speed situations. the potential to reduce this backpressure is higher. Figure 8: Mach number for Original and DEP concept engine at 6500RPM 6 CONCLUSIONS The general trend under fixed BMEP condition is that at lower engine speed. It is attributed to the PMEP reductions and more importantly due to the smaller inertia of the DEP turbocharger. However. We also wish to appreciate the China Scholarship Council to partly sponsor my living expenses. the backpressure is low and there is less to gain from the DEP concept whilst at higher speeds the backpressure is higher. introducing fully flexible intake and exhaust valve mechanism could further optimize the results above. Using the DEP concept for entire speed range due to the better cylinder scavenging can increase BMEP. Transient performance for the DEP concept engine is considered to be better than the baseline engine. The DEP-based engine can achieve improved BSFC and BMEP performance due to better gas exchange and combustion processes without changing the settings of the intake and exhaust valve profile. Thus the BSFC reduction is more noticeable at higher speeds. the trend for RGF under high and low speed range is different. 38 . Valve-Event Modulated Boost System. Stockholm. H. “Four-Octane-Number Method for Predicting the Anti-Knock Behavior of Fuels and Engines”. Relative Air-fuel Ratio. (2012). 39 . [11] Hong.. Blow-Down and Scavenging Valve Area”. C. [7] Roth. SAE International. Divided Exhaust Period . D.8 REFERENCE LIST [1] Möller. “Valve-Event Modulated Boost System: Fuel consumption and Performance Potential”. and Load on SI Engine Efficiency. F. [5] Gundmalm. Engines 5(2): 538-546. “Divided Exhaust Period: Effects of Changing the Relation between Intake.. SAE Int. J.A Gas Exchange System for Turbocharged SI Engines. (2005). (2010).. 72. et al. and M. P. (2013). [4] David Roth. Engines 6(2). 1978. D. [6] Gundmalm. “Valve-Event Modulated Boost System: Fuel Consumption and Performance with Scavenge-Sourced EGR”. SAE Int. S. Becker (2012). et al. D. [2] Westin. (2000). KTH Royal Institute of Technology: vi. A. SAE International. J. (2013). Review and analysis of variable valve timing strategies . S. Divided Exhaust Period on Heavy-Duty Diesel Engines. Compression Ratio. et al. B. [10] Douaud. Professional Engineering Publishing. F. B. doi:10. Effects of Combustion Phasing. [3] Ayala. P. D. B. and Eyzat.. [9] Gamma Technology. SAE International. M. E..Eight ways to approach. et al. The Influence of Residual Gases on Knock in Turbocharged SI-Engines. [8] Roth. (2004). SAE Technical Paper 780080. A. (2006). SAE International.. et al..4271/780080. et al. E Binder. the supercharging system provides even further potential for increasing mean effective pressure by 10% at same low end torque speed. Germany ABSTRACT This paper examines the influence of future boundary conditions for boosting systems. It employs engine process simulation to select different intake valve timings on a four-cylinder 1. NOMENCLATURE BMEP Break mean effective pressure n speed in rpm pcyl Cylinder pressure in bar m mass flow kg s PFI performance index P Power in kW Greek Letters η efficiency Π pressure ratio Subscripts C compressor d displacement exh exhaust F friction Rel related Rel related red reduced RP rated power is isentropic s static Spec specific T turbine TC turbocharger LET low end torque measured measured value m mechanical max maximum WG wastegate _______________________________________ © The author(s) and/or their employer(s). package and catalyst light-off. Measurements on a single-cylinder engine were able to confirm the fuel-saving trends identified in simulation. wastegate and VTC. Despite demanding a higher level of boost pressure.4-l spark-ignition engine with a view to illustrating the extent to which a single-stage boosting system is capable of providing the level of boost pressure that is demanded.Advanced boosting technologies for future SI engine concepts P Grigoriadis. the most promising configuration makes it possible to attain the basic full-load curve using Miller timings (∆IVC=40°). Single-stage boosting systems show a number of benefits over two-stage systems in terms of cost. Comprising VTG. This was the reason for examining how far adapting or advancing a single-stage boosting system can meet the rising demands on boost pressure. 2014 41 . L Böttcher. Variabilities on the compressor and turbine side were used to increase boosting system spread. This reduces consumption in the region of 3 to 6%. M Sens IAV GmbH. these being increasingly influenced by exhaust gas turbochargers. This is why discussion today is looking closer at technologies that can produce a positive effect on the entire engine operating range. If applicable. are adjusted until they reach the target set. the range considered will extend to the entire engine operating range. the question must be asked as to whether aerodynamics can deliver further potential at reasonable cost. made up of a rigid radial compressor and a rigid radial turbine with wastegate. new methods are briefly explained in the next section.2 Advanced Turbocharger Model The advanced turbocharger model basically comprises the following elements: . This means that engine concepts undergo preliminary evaluation on the basis of simulation results. Aerodynamics of compressor and turbine on the basis of maps obtained from experiments . 2. has the single-stage boosting system reached its limit? Engine process simulation is a tool that is suitable for examining a question of this type. is seen as a standard. Transient engine behavior is understood to mean both an engine’s responsiveness (time to torque) as well as catalyst light-off. with transient engine behavior carrying greater weight in the evaluation process. However. A key aspect in this process is the advanced turbocharger model that is explained in more detail in the next section. Friction on the basis of a map obtained from experiments . emissions and response behavior while keeping a focus on the applicable exhaust emission testing and consumption cycles. In recent years aerodynamic measures have made it possible to increase throughput spread and efficiency of such radial machines. promising parameters. With exhaust- gas turbocharged SI engines. The simulation results are examined to establish whether they reach target values. the boosting system must be adapted to accommodate the changed engine boundary conditions.1 Modern Development Process The modern development process integrates 1-D engine process simulation to an even greater extent into configuring an engine concept. Figure 1 shows the general path taken in a modern development process. such as the aerodynamic behavior of a turbocharger. fuel consumption and responsiveness. For this reason. 2 METHODS AND TOOLS IN THE MODERN DEVELOPMENT PROCESS 2. For this purpose and as a technology element now firmly established in engine development. These initially synthetic and iteratively obtained maps then provide the basis either for selecting an appropriate turbocharger or for configuring new rotors. The latter can be achieved in the broadest sense by means of downsizing. In current cycles the time-averaged engine operating range tends to concentrate on the part-load section. such as drivability at full load.1 INTRODUCTION Engine developers move between the poles of fuel consumption. downspeeding and lean combustion processes. The methods for reproducing a boosting system in engine process simulation have been reconsidered in recent years. If not. Heat flow and heat storage capability on the basis of semi-empirical approaches 42 . the focus is on avoiding full-load enrichment and minimizing throttling losses. For future demands. Single-stage boosting. such as Real Driving Emissions (RDE). The latter has already been used with success in [2]. 43 . In essence. Diabatic criterion χ to assess heat transfer behavior in qualitative terms . Exhaust gas flap for increasing turbine mass flow to reduce the heat transfer . the turbocharger is operated in a transient state both on the real-world test bench as well as in a model test rig. To validate the model. A further measurement on the turbocharger test bench under normal “hot” boundary conditions permits validation of the heat model. Semi-empirical approaches are then used for extrapolating map ranges that cannot be measured. however. 2) PT can be used for determining frictional power PF or the level of mechanical efficiency ηm. 1) PC m (Eq. The heat transfer normally occurring are avoided as far as possible here by adapting the measurement boundary conditions. The procedure required for this is described in detail in [1]. PF PT PC and (Eq. Closed-loop system for extending turbine maps and for increasing compressor mass flow to reduce the flow of heat . The heat-storage capability of a turbocharger is also simulated and is essentially based on turbocharger mass. Calculation of turbine power output PT by use of temperature upstream and downstream of the turbine Proceeding from the power outputs available for compressor and turbine. The range to the left of the surge limit and the range in the fourth quadrant of a compressor map are of particular relevance to acceleration cycles and to the low-end torque operating point (LET). Figure 1: Modern development process for configuring a boosting system for supercharged engines Aerodynamic behavior is ascertained directly on IAV’s turbocharger test bench. Details on this are provided in [3]. it is based on the following methods and assumptions: . The model’s quality is then determined on the basis of how well the simulated and measured transient turbocharger speed curve match up. 1 Torque Spread. Combustion process limits. Power Density and Responsiveness Usually. Furthermore.2. the demands on an engine are such that maximum attainable engine torque is even reached at relatively low engine speeds. The relevant characteristic value is calculated as follows PRP PRP. such as engine knock. 80 kW/l. the underlying. Figure 2: Schematic diagram of the interaction between turbocharger and engine in the turbocharger nomogram 3 CURRENT AND FUTURE BOUNDARY CONDITIONS FOR BOOSTING SYSTEMS 3.3 Turbocharger Matching Classic turbocharger matching is based on the ability to place an engine’s target full-load curve into a compressor map in such a way as to ensure a safety margin to the compressor’s surge limit and throughput limit. In future. but can also reach peak levels of up to 133 kW/l [4]. compressor output and turbine output are to be equal at each steady-state operating point. Spec (Eq. For a better understanding.Re l (Eq. simplified engine air mass flow curves do not take account of any valve overlap or recirculated exhaust gas. IAV has defined a characteristic value that can be calculated as follows BMEPLET 1000 BMEPLET. 4) Vd 44 . will also be included (see Figure 2). allowance will be made for engine boundary conditions of this nature. As a rule. 3) nLET Over recent years power density has constantly increased and averages at approx. 35 rising to a maximum of 0. the naturally aspirated engine component of an engine map becomes increasingly smaller as power density rises.Spec) are given by the maximum known values in the database.Spec BMEPLET.Both characteristic values can be combined to produce a relative performance index PFIRel which is shown in Figure 3 in the form of an isoline. This means that a boosting system has more and more of an influence on any requested sudden load increase from part to full load. the impact of the mass moment of inertia associated with the turbocharger’s response becomes increasingly pronounced.Re l PFIRel (Eq.max BMEPLET. As a result. diminishing the responsiveness of an engine with high power density on account of the design principle involved.7. 5) PRP. data taken from IAV’s Engine Knowledge Database 3. PRP. retarding it as Atkinson timing. Apart from dethrottling the engine.2 Effective Engine Displacement Throttling losses can be reduced by using variabilities in the valvetrain on the basis of various technological approaches. The basis for these data is taken from IAV’s Engine Knowledge Database. that the maximum values for the attainable torque at low engine speeds (BMEPLET. The point at which the intake valves close is selected in a way that appears to reduce an engine’s displacement. This has a positive effect on the engine’s knock 45 .Spec.max On top of this.Re l. ultimately diminishing throttling losses. producing a reduction in the final compression temperature. Today’s volume-produced engines come with an average relative performance index of 0. It has to be mentioned.Rel) and the the power density (PRP. the cylinder charge is also relaxed with Miller timing. Advancing the closing cycle is usually referred to as Miller timing. Figure 3: Relative performance index PFIRel for different engines. Engine power output is prevented from falling by opening the throttle. in principle following the procedure described in Section 2. which means that the turbine can ultimately be provided with less exhaust gas energy.characteristics at high loads. is then globally raised to as much as 4-5. For the same engine power output. have already been described [5].1 Basic Engine Model Engine process simulation provided the basis for examining engine behavior. 3. 3.4 Number of Cylinders Reducing engine displacement while maintaining effective engine output is known as downsizing and. Any endeavor to reduce effective engine displacement across the entire engine operating range for the above-mentioned reasons demands a higher level of boost pressure – particularly above natural aspirated full load – which must then come from the boosting system.3 Lean Combustion Process The lean combustion process involves providing an air-fuel ratio in the combustion chamber that is leaner than stoichiometric. among other aspects. Lacking turbulence lengthens the inflammation phase and significantly increases the duration of combustion. The decrease in airflow rate also fluctuates widely on the fresh-air side. ignition timing is automatically retarded. ignition timing can be moved into ranges that provide greater efficiency. Added to this. If knock probability increases. these two values produce a relative performance index of 0. Countermeasures to oppose this.4 l four-cylinder spark-ignition engine with an effective power output of 96 kW and an LET of 19 bar. Because of the higher airflow rate required. ultimately producing relatively low levels of turbine efficiency. As a result.36. and the engine model provides predictive combustion development computation – SI Turb – as well as a knock model. gains its advantage from shifting the operating point in an engine’s operating map. particularly if lean-burn operation is also to be used above the naturally aspirated engine load range. In the high engine power output range. producing a fall in throttling losses. a lean air-fuel ratio leads to lower exhaust gas temperature. this can result in unstable compressor operation and consequently limit achievable engine load [9]. this can lead to turbine choking. 46 . Applied in equation 5. Particularly in surge-limit vicinity. particularly if reducing displacement involves a smaller number of cylinders [7]. This means that a highly fluctuating supply of energy is admitted to a turbine. the value of 1. usual for spark-ignition engines. The lean combustion process requires a higher level of ignition energy and places elevated demands on the exhaust gas aftertreatment systems on account of the mixture being lean. Boost pressure is regulated by a wastegate. Miller timing tends to impede the formation of intake-port-induced charge motion in the cylinder which dissipates into turbulent kinetic energy (TKE) in the further course of piston movement. the mass flow characteristics in the exhaust gas and air path with high time resolution then exhibit higher amplitudes. 4 INTERACTION BETWEEN ENGINE AND BOOSTING SYSTEM 4. The engine model used is an exhaust-gas turbocharged 1. The main center of heat release is set to produce maximum efficiency (8 °CA ATDC). the lean combustion process also demands a higher level of boost pressure from the boosting system [6]. producing a negative influence on combustion. in the region of natural aspirated full load. [8]. however. The valvetrain provides variabilities at the intake. The simulation environment is GT-Power. Depending on the lean concept implemented. Downsizing is also used for reducing friction losses. 3 Selecting and Adapting the Boosting System To set Miller timings without having to accept the above-described losses in engine load. Additional mass in the exhaust gas path (2nd turbine) acts as a heat sink. the turbocharger is pushed to its speed limit. At 5. however.500 rpm for the basis configuration with Miller timing (filled circular symbol). further adjustments must be made to the boosting system. The basic turbocharger was then scaled down in size as shown in Figure 5 (rectangular symbols). The drop in torque is attributable to the reduced engine charge from advanced “intake closes”.000 rpm. Despite Miller timing. Proceeding from basic timing. Although a serial-sequential two-stage boosting system consisting of two exhaust gas turbochargers would deliver this spread [10]. with the full-load curve once again being determined for the engine. this section examines the influence of Miller timing (see Section 3.4.4 bar. higher boost pressure is required across the entire rpm range. High cost 47 . equating to a power loss of 32kW.7 bar to 9. this involves no changes to the basic boosting system. particularly in the heating phase. it is now possible to achieve the basic full-load curve again.2 Boost Pressure Demanded from Changed Engine Boundary Conditions By way of example. In principle. Figure 4: Full-load curves for the 1. Initially. As from a speed of 2. High demands are placed on ensuring a tight seal at the exhaust gas flap additionally installed in the exhaust gas path as it needs to switch between high-pressure and low-pressure turbine . adversely affecting catalyst light-off . it cannot be considered for the present on the following grounds: .4-l spark-ignition engine from different valve-train strategies and for different boosting systems 4.500 rpm. this being accompanied by higher boosting system spread. mean effective pressure falls from 14. Figure 4 shows the full-load characteristics for the basic configuration (unfilled circle symbol) and torque reduced by 30% at 1. High demands on packaging . the intake valve timing was advanced by 40° CA.2) on the demand for boost pressure. a two-stage serial-sequential system could be restricted to the fresh-air side alone. ultimately. However. packaging and. relatively high cost. less potential to save fuel all work against the electrical system. This is done by fitting both compressor and turbine with a variable guide vane system. High load on the vehicle electrical system. it should be remembered that in the engine concept considered here an additional compressor of this kind must be in a position to deliver boost pressure over a prolonged period. The last three points also apply to the additional mechanical system. Figure 5: Turbocharger maps used in simulation for producing different configurations However. The basic compressor is equipped with a VTC system (Variable Trim Compressor) that can alter the effective inlet cross 48 . For the above reasons it would make sense to get the single-stage boosting system to meet the underlying demands. Combinations consisting of an exhaust gas turbocharger compressor and an electrically or mechanically driven compressor are conceivable. Although a VTG turbine has also been used in the SI engine segment since 2005 [12]. however. This is because it was possible on average to advance the main center of heat release by 3 °CA and reduce pumping work by an average of 130 mbar. On the hot gas side. As this is a variable system. the trim appears capable of changing the ratio between rotor inlet and rotor outlet diameter. As shown in Figure 4. VTG turbine). this must be made larger than the rotor of a wastegate turbine. it is still a niche application.000 rpm even 490 mbar. this configuration is referred to as “single-stage advanced+”. is put at about 30%. On account of its throughput spread and relatively high levels of efficiency. enabled by the VTC system. the engine operating points are within the stable map range. however. comes with relatively minor loss in turbine efficiency. 300 mbar. For the engine under study here. is open all the way. this influence is relatively slight which probably has to do with the basic configuration’s valve overlap being low to start with. The latter were synthesized on the basis of experience gathered from experiments and by employing analytical approaches. This. thereby resulting in a shift in the surge limit (up to 33%). This corresponds to an increase of 10 %. As the entire exhaust gas mass flow has to be fed through the rotor of a VTG turbine. the turbine is able to provide the compressor with the necessary drive power. Using the supercharging unit’s full potential to increase power output does. at 4. Proceeding from the LET operating point. an attempt was made to achieve the same mean effective pressure. the VTG system continually opens and. In addition to the previously mentioned turbocharger maps for the basic (rigid compressor. As a result.9 to 21. This measure ultimately makes it possible to achieve the basic full load curve (see Figure 4. 49 . taking the relative performance index to 0. the turbine pressure ratio and. rotor diameter is therefore increased from 37 to 40 mm. the exhaust gas backpressure. In the studies conducted here. Reducing the inlet cross section by 40% shifts the map towards smaller mass flows. with this. At the LET the boost pressure differential is approx. IAV’s own matching tool was used to match the VTG turbine to the engine in such a way as to achieve the engine’s LET operating point.section of a rotor [11]. diamond symbols). the VTG turbine provides a wider spread in the turbine map range. At maximum engine power output. As the compressor map is wider for the “single-stage advanced+” configuration.500 rpm with throttle position unchanged (100%).44. This geometric modification is always accompanied by an increase in the mass moment of inertia and. the initial cross section can be reproduced without any additional pressure loss in the air path. As the VTG system cannot be opened any further. the turbocharger reaches its speed limit above 3. on the basis of available data. wastegate turbine) and the scaled turbocharger Figure 5 shows the turbocharger maps for the advanced boosting system (VTC compressor. Specific fuel consumption can be lowered by between 3 and 6% along the engine full-load curve. This necessitates further increasing turbine throughput spread and turbine efficiency which is done by adding a wastegate flap. the wastegate turbine currently used is replaced with a VTG turbine familiar from diesel engines. The compressor map shows the higher level of boost pressure required across the full-load curve. at an engine speed of 3.500 rpm. In just the same way as the wastegate turbine. The nomogram in Figure 6 presents the full-load curves for the “basic configuration” and “single-stage advanced+” options. Below. are shown to remain at about the same level in spite of boost pressure being higher. the basic full-load curve cannot be completely reproduced with this configuration (triangular symbol).9 bar and maximum effective power output from 96 to 107kW. In the investigations presented hitherto. make it possible to increase maximum mean effective pressure from an initial 19. The extent to which avoiding scavenging can influence the boosting system was also examined. 50 . Proceeding from an almost closed wastegate and open VTG system at standard timings. Figure 7 shows the savings measured in specific fuel consumption. however. The white rectangle represents a simulated engine operating point with a turbine downsized by 25%.4% (green rectangles) by increasing Miller timing. In general. However. This demonstrates the benefits as well as the limitations of additional variabilities. “intake closes” timing was gradually advanced while adjusting boost pressure and taking the air-fuel ratio to 1.1. be needed to ascertain whether it is possible to achieve the effective power output of 96 kW with this turbine. experiment-based studies were conducted on a single-cylinder research engine already used in [5]. This has a fully variable valve train and provides the capability of giving the injector a centered or lateral position. an increasing scavenging–pressure differential as well as increasing Miller timing are both seen to have a positive influence on specific fuel consumption.500 rpm. closing the VTG by any more takes the improvement in consumption back to 4. At this operating point. Further investigation would. In a test series at constant load (n=1. Figure 6: Turbocharger nomogram for the basic configuration (black) and the "1-stage advanced+" configuration (red) 5 EXPERIMENTS ON THE SINGLE-CYLINDER ENGINE To validate the above-described simulations. specific fuel consumption can be reduced by approx. 5. Further simulations were carried out using the engine model described in 4.3%. consumption can only be improved further by reducing turbine size. full load). Advanced Turbocharger Model for 1D ICE Simulation ... The potential is also given for satisfying the other future boundary conditions stated above.. Figure 7: Fuel savings measured and computed while varying scavenging-pressure differential and timing at operating point n=1. doi:10. R.500 rpm and full load 6 SUMMARY AND OUTLOOK It was possible to show that changes in the engine’s boundary conditions have a considerable impact on engine behavior. Sens. Benz. Mean effective pressure can be increased by 10% as well. Method of performance measurement for low turbocharger speeds. the higher demand for boost pressure can also be met for engine operating points between intake full load and full load.4-l spark-ignition engine as the basis for examining the behavior of different single-stage boosting systems. M. London 51 . This configuration is capable of achieving the basic full-load curve.. Berndt. SAE Technical Paper 2013-01- 0581. 7 REFERENCES [1] Grigoriadis. E... On account of high throughput spread. the “single-stage advanced+” configuration was shown to be highly promising in this regard. Böttcher. et al. such as the lean combustion process. Miller timing was applied to a 1. P. 9th International Conference on Turbochargers and Turbocharging. Measurements on a single-cylinder engine were able to confirm the fuel-saving trends identified in simulation. L. By way of example and representative of future technologies. 2010. Binder. T. A. P. IMechE. Comprising a variable trim compressor (VTC) and a variable geometry turbine (VTG) with wastegate..4271/2013-01-0581 [2] Otobe.Part I.. Grigoriadis. 2013. J. 11th Supercharging Conference. Micelli. C.. et al. M. Article in MTZ worldwide Edition: 2013-07 [6] Kneifel. H.. Vienna [5] Riess. Lingenauber. Müller. 2011. Intake Valve Lift Strategies for Turbulence Generation. Kiener. P. S. 22nd Aachen Colloquium Automobile and Engine Technology. Sens. Innovative Two-Stage Turbocharging System with Cooled Regulating Valve for Gasoline Engines. Aachen [9] Cuniberti... Friedfeldt.. 16th Supercharging Conference.. SAE Technical Paper 2008-01-0143. R. D. Charging system for a small bi-cylinder engine: the TwinAir experience.. 34th International Vienna Motor Symposium.4271/2008-01-0143 [7] Lee... A. IMechE.. S. R. Sens. A. Wöbke..[3] Grigoriadis.0 l Three- cylinder Gasoline Engine. Der Turbolader mit variabler Turbinengeometrie (VTG) für den neuen Porsche 11 Turbo – Ein Meilenstein in der Ottomotorenaufladung. Ph. Thesis.. Benz. Stroppiana. K. 2012.... Becker.. Sens. doi: 10. Venezia. Experimental data acquisition and modeling of unsteady flow phenomena of vehicle engine turbocharger compressors. M.. The New 3 Cylinder 1. 10th International Conference on Turbochargers and Turbocharging. Investigations on Supercharging Stratified Part Load in a Spray-Guided DI SI Engine. Kim. 2013.... 2013. J. M. Werner. et al. M. M. 2006. M. Gu.D.... 2013. 2008. Variable Trim Compressor – A New Approach to Variable Compressor Geometry. T. T.. 2011.. Dresden [10] Kuhlbach.. M. P. Der neue Hochleistungsvierzylindermotor mit Turboaufladung von AMG. Aachen [11] Grigoriadis. Pape.. 20th Aachen Colloquium Automobile and Engine Technology. S. Article in MTZ worldwide Edition: 2011-08 [8] Ernst.. Germany [4] Hart. J. T. London [12] Gabriel. Technical University Berlin. 2008.0L Gasoline Direct Injection Turbo Engine from Ford. Dresden 52 . R.. Lamb. Ramb. 2008. 2011.. M. A. Y.. The New Hyundai-Kia 1. Hahn. Japan ABSTRACT The design of a mixed flow compressor stage with an extremely high flow coefficient ( ) of 0. IHI Corporation. IHI Corporation.25 and a high pressure rise coefficient ( ) of 0. Japan H Tamaki Corporate Research and Development.56 is described. Germany R Numakura Turbo Machinery and Engine Technology Department. NOMENCLATURE Cm meridional velocity (m/s) Cp isobaric specific heat capacity (J/kg K) diameter (m) total enthalpy (J/kg) absolute Mach number (-) = / tip speed Mach number (-) mass flow rate (kg/s) SM Surge Margin (-) temperature (K) blade speed (m/s) = / volume flow rate based on inlet total density(m3/s) relative velocity (m/s) absolute flow angle from radial direction (º) relative flow angle from radial direction(º) total-to-total efficiency (-) pressure ratio (-) density (kg/m3) ___________________________________________ © The author(s) and/or their employer(s). The objective of the work was to explore the performance potential in this highly unconventional area of the design space and to assess the capability of design methods. UK M V Casey PCA Engineers Ltd. University of Stuttgart. The paper describes the aero-mechanical design approach for the preliminary design and discusses the challenges involved in developing such highly loaded compact stages. UK Institute of Thermal Turbomachinery (ITSM). The test data obtained on a prototype stage is also presented.Design and testing of a high flow coefficient mixed flow impeller H R Hazby PCA Engineers Ltd. The results show that acceptable performance levels can be achieved at these extreme design conditions and further exploration of the design space is worthwhile. 2014 55 . 02 m3/s and a rotational speed of 77525 rpm. Mixed flow stages also have lower radius change across the impeller than stages with lower flow coefficients so that a high pressure rise is more difficult to achieve than in conventional stages as there is a smaller centrifugal effect. The impeller outlet diameter was fixed at 100mm. but to what extent this is possible in the relatively uncharted territory of mixed flow designs is not known. Some guidelines for the expected performance and the preliminary design of radial stages with a high swallowing capacity have been published by Rusch and Casey (2).56. the potential performance levels and operating range that can be achieved by high flow coefficient mixed flow impellers have not been thoroughly addressed in the literature. Detailed design of the stage and further analysis of the performance can be found in Hazby et al. the specifications of the final compressor design accompanied with the performance measurements are presented.25 and an isentropic pressure rise coefficient of ( = ∆ / ) of 0. (1) can be used to place the design point in the recommended design space for centrifugal compressors. the guidelines of Casey et al. The requirement for compact centrifugal compressors with high swallowing capacity leads to the application of small mixed flow impellers with extremely high flow coefficients (1). The small size and high speed increases the Mach numbers in the flow channels and this may conflict with the achievement of good efficiency and wide operating range. This paper attempts to contribute in this area by discussing the issues faced in the design of mixed flow impellers for extreme duties and the achievable performance levels. Furthermore. 56 . As a first indication of the difficulty of the design. The target design point for the stage was set at a flow coefficient ( = / ) of 0. there is a strong requirement for compact designs to reduce the size and cost of the installation. which is far outside of the conventional design space for centrifugal compressors. such as diesel engines for standby power. based on measurements on a prototype compressor stage. as a function of the design flow coefficient and the required isentropic pressure rise coefficient. clear design guidelines for this type of turbomachine have not been published. (3) 2 CONCEPTUAL DESIGN CONSIDERATIONS The current impeller was designed to deliver a pressure ratio of 2.65 at an inlet volume flow of 1. In the next sections the general conceptual design issues regarding the design of very high flow coefficient mixed flow impellers are discussed and the impacts of the mechanical integrity requirements on the aerodynamic design of the compressor are addressed. However. The use of modern design methods is expected to alleviate the performance deficit arising from the extreme duty.Subscripts 0 stagnation properties 1 impeller inlet 2 impeller outlet m mean diameter h hub diameter s isentropic flow process t tip diameter 1 INTRODUCTION In many turbocharger applications. Then. defined below. Hence. For comparison. For a constant rotational speed and pressure ratio the pressure rise coefficient decreases as the outlet diameter is increased whereby the design with a 120mm tip diameter is placed within the recommended design range. 0.1 0.2 0. = (1) ∆ = = (2) Guidelines for the achievable pressure rise coefficient at a particular design flow coefficient are given in Figure 1. facing the designer with a number of challenges regarding the aerodynamic performance and mechanical integrity of the wheel. The pressure rise coefficient for the current impeller at the design operating point is 0. the impeller was designed at a mean outlet diameter of 100mm.15 0.1 0 0 0.35 Flow Coefficient Figure 1 Comparison of the effect of the impeller diameter on the location of the design point in a modified Cordier diagram.56. The dashed lines represent the range of the current experience. The aero-mechanical design considerations involved in the development of such stages is discussed in more detail in the next sections.25 0.4 120 mm 0. If the rotational speed is maintained constant. the pressure rise coefficients corresponding to impeller designs with outlet diameters of 110 and 120 millimeters are also shown in the figure. The intention of this study was however.3 0. 57 . A similar diagram can be found in Dixon and Hall (4).05 0.5 110 mm 0.3 0. which is a modified form of the Cordier diagrams given in (1). the reduction in outlet diameter can drive the impeller flow coefficient and the pressure rise coefficient to extremely high values. which is well outside of the range of the conventional centrifugal compressor design shown in Figure 1.6 Pressure rise Coefficient 100 mm 0. to push the limits of conventional designs as far as possible into the mixed flow region and assess the achievable performance levels in this design space. compact centrifugal compressor stages can be produced by reducing the impeller outlet diameter.2 0. The general trend is a continuous reduction in the pressure rise capacity of the impeller as the flow capacity is increased. For the same compressor duty in terms of pressure ratio and mass flow rate.7 0. 07.52 / = 0.93 / = 0. a higher curvature is needed on the shroud to turn the flow from axial to radial direction.25 flow coefficient radial impeller.86 for all cases.25 = 0.89 / = 0. It is the variation of these parameters that influences the design constraints and performance of the impeller as its size is reduced. shown in Figure 3 and Figure 4.72 = 0. It should be noted that the extreme flow acceleration observed near the outlet of the 0. corresponding to an outlet diameter of 100mm.22 / = 1.47 = 0. This results in the acceleration of the high speed flow on the shroud and increased hub-to-shroud secondary flows inside the passage. The flow acceleration on the shroud can be seen in the results of the throughflow calculations.17 / = 1.69 / = 0. shown in Figure 2. These effects are discussed with the help of Figure 2. using the throughflow code Vista TF as published by Casey and Robinson (3).61 / = 1.05 / = 0. A difficulty that arises from reducing the outlet diameter is that with the inlet tip diameter fixed.89 = 0. with an impeller diameter of 120mm has the lowest flow coefficient of 0.56 = 54° = 40° = 18° = 18° = 55° = 55° = 55° = 55° = 0. The impellers were designed for the compressor duty as mentioned above and assuming a constant impeller total-to-total efficiency of 0.25 = 0.59 / = 0.15 = 0. 58 . In this impeller.07 / = 1. designed for the same duty but with different outlet diameters.39 = 0.09) due to relatively high inducer losses (see Rusch and Casey (2)).80 = 0.16 Figure 2 Effect of outlet diameter on the impeller design at the same duty of pressure ratio and mass flow rate.15 which is considered above the value for an optimal efficiency (which is about 0. relative diffusion and the backsweep at the outlet of the impeller. The required pressure ratio was achieved by varying the backsweep angle as the impeller tip diameter was reduced.56 = 0. = 120 = 110 = 100 = 100 = 0. does not happen in reality as the flow will most probably separate on the sharp corner. The tip diameter ratio of / = 1. which has a tip diameter ratio of 1. In this section some more detail is provided of the geometry of possible competing stages for this application and the relevant aerodynamic design issues are discussed.80 / = 1.28 suggests a relatively sharp turning of the flow on the shroud contour of this impeller. The amount of pressure rise that can be achieved at a particular mass flow rate is primarily a function of the wheel diameter. the absolute flow angle at the inlet of the diffuser is kept equal to 55˚ for all cases. which demonstrates four impellers. turning of the flow from axial to radial is not practically possible due to the high curvature near the impeller outlet and therefore a mixed flow design with the same flow coefficient becomes necessary.3 PRELIMINARY DESIGN CONSIDERATIONS The conceptual considerations given above identify the level of difficulty in terms of the global experience with centrifugal compressors from other sources. To obtain a fair comparison. The first stage.19 = 0.28 / = 1. The higher inlet relative velocity at the tip of an impeller blade means that the flow streamline at the tip section undergoes a stronger diffusion from inlet to outlet than that on the hub. However. where the variations of the mean relative Mach number is shown for hub and shroud contours for all four impellers.20 100mm mixed flow 0.8 1.4 0.30 0.60 Mean relative Mach number Mean relative Mach number 0.80 1.40 0.8 1.20 0.00 100mm mixed flow 0.60 110mm radial flow 0. This can also be seen in Figure 4.93 when the impeller outlet diameter is reduced from 120mm to 100mm. The accelerated flow on the hub is generally more tolerant to the angle variations.40 100mm radial flow 0.00 0.2 0. the ratio of the outlet to inlet relative velocity at the tip section is close to 0. the deceleration at the hub increases as the outlet diameter is reduced.40 0.0 0.20 0.15 = 0.19 = 0.25 = 0. In a typical centrifugal compressor. about 50% of the pressure rise is generated by the centrifugal effects due to the rotation of the fluid particles about the axis.2 0. relieves the diffusion on the casing but further increases the diffusion on the hub by more than 15% due to the reduction in the hub diameter at the outlet. which means more diffusion needs to be done inside the passage to achieve the same level of pressure rise. This is an efficient means of compression as it is independent of diffusion process and flow separations inside the narrow passages. However.00 0. In a typical centrifugal impeller such as stage 1 in Figure 2.70 120mm radial flow 1.0 0. In the limiting case of an axial compressor it is actually the hub section which imposes the critical limit on the loading.6 0.80 120mm radial flow 0. the flow on the hub section in fact accelerates due to a high relative velocity at the impeller outlet compared to that at the inlet. This makes the diffusion on the shroud a critical constraint for the impeller design with values of relative velocity ratio of greater than 0. This effect is reduced as the flow coefficient is increased (outlet diameter is reduced).50 100mm radial flow 1. 59 .0 Meridional distance Meridional distance Figure 4 Variation of the mean relative Mach number on the hub and shroud contours.4 0.10 0. due to low blade speed in the hub section. The higher overall diffusion level that is needed at smaller outlet diameters is partly due to the smaller centrifugal pressure rise inside the impeller passage. Moving from a radial design to a mixed flow design with the same mean outlet diameter and 45˚ meridional exit angle.60 110mm radial flow 1.22 to 0.7.25 Figure 3 Meridional velocity contours from a throughflow calculation of the impellers shown in Figure 2.5 usually used to avoid excessive diffusion at the tip section.6 0. Cm (m/s) = 0. Hub section Tip section 0.0 0. This can be observed in the value of / given in Figure 2 where it reduces from 1. 110 and 100 mm diameter. Figure 5 Expected variable speed performance maps for the three radial compressors with 120. designed for the same aerodynamic duty. This can be observed in the expected performance maps of the three radial stages described above with 120mm. This is particularly true for high speed impellers where the control of the supersonic flow in the inducer region has a significant impact on the performance and operating range of the compressors. shown in Figure 5. modified to be used for very high flow coefficient impellers similar to that of the current study. 60 . opposes the requirement of higher relative diffusion at the hub of the high flow coefficient impellers. demanding a greater compromise on the thickness and camber distribution at the hub section. as predicted by the method of Casey and Robinson (5). the aerodynamic design of the tip section of a centrifugal impeller is more critical than the design of the hub section due to its higher diffusion levels. 4 MECHANICAL DESIGN IMPLICATIONS As discussed earlier. A validation of the code for mixed flow stage is provided in Casey et al. The natural frequency of the impeller vanes decreases significantly as the outlet diameter is reduced. The position of the target design point is depicted by the red circles in the pictures. The performance maps were predicted using the method presented by Casey and Robinson (6). The required low backsweep angle results in narrower characteristics with smaller surge margin by moving the peak efficiency operating point closer to the surge line. 110mm and 100 mm diameters. however.The lower centrifugal pressure rise due to smaller impeller outlet diameters cannot be entirely compensated by a reduction in the relative velocity ratio and therefore smaller backsweep angles are needed to deliver the target pressure ratio. which could limit the freedom of the design in this section. The diagrams identify that considerable difficulty will be expected with a small operating range as the characteristics become less steep as the backsweep is reduced for the smaller stages. As can be seen in Figure 2. A higher backsweep angle was not used for the 100mm diameter design as this would increase the relative diffusion at the impeller tip to unacceptably high levels. Therefore. a greater degree of freedom is often exercised in the design of the hub section to satisfy the mechanical integrity criteria. the required backsweep angle decreases from 54º to 18º as the impeller outlet diameter is reduced from 120mm to 100mm. This. (7). The impeller geometry and meridional shape of the passage are presented in Figure 7.25. Further details of the mechanical design are given by Hazby et al. which was required as a design constraint.02 m3/s and a rotational speed of 77525 rpm. The impeller vane profiles were designed at several spanwise locations to optimize the aerodynamic performance of the impeller by controlling the supersonic flow at the inlet as well as ensuring the mechanical integrity of the wheel. optimized for handling the high speed flow. (3). High stress region Figure 6 Stress distribution in the main and splitter vanes. it requires thinner splitter vanes with their leading edge placed further downstream in the passage to allow a sufficient flow capacity at the hub section. The final impeller consisted of 9 main vanes and 9 splitters and was designed with a back sweep angle of 28˚ degrees at the outlet mean radius. As discussed in section 3.These mechanical requirements also compromise the aerodynamic design at higher spanwise positions. 5 FINAL STAGE DESIGN AND TEST DATA In the previous sections. The outlet mean diameter was set at 100mm.65 at an inlet volume flow rate of 1. corresponding to an extremely high flow coefficient of 0. the choice of the vane thickness can limit the number of the vanes that can be fitted on the hub section. Under these circumstances the aerodynamic optimization of the splitter vanes is very much limited by the mechanical design requirements. (3). Further details of the aerodynamic design of the blade sections are given by Hazby et al. This in turn can alter the detailed design of the tip section. a mean outlet diameter of 100mm requires a mixed flow style impeller. This meant that the conventional straight line generators needed for flank milling of the impeller blades could not be used and therefore the impeller was point milled to allow the use of curved line generators. This impeller was designed to achieve a pressure ratio of 2. various aspects of the aero-mechanical design of a high flow coefficient centrifugal impeller were discussed. 61 . The final design of the stage is presented here. In addition. At the exit of the impeller. the flow was turned to the radial direction using a converging vaneless diffuser followed by a parallel walled diffuser section. where the vane thickness and curvature need to be minimized to reduce the aerodynamic shock losses. where turning of the flow from axial to radial is partly accomplished in the vaneless diffuser. The result could be swept splitter vanes with high stress levels in the vane area as shown in Figure 6. Furthermore. The constraint was applied only on the outlet mean diameter and therefore the hub and tip diameters could change as the meridional position of the blade at the trailing edge was not restricted. the coefficients in the 1D prediction method have been adapted to match the current stage performance. Figure 7 Meridional passage and impeller geometry. stage pressure ratio versus volume flow. are also shown in the figure. Overall. The stage as designed was tested in a dedicated turbocharger test stand. 62 . At near design speed of 77100 rpm. The surge margin at a nominal pressure ratio can be defined as: SM = (Design flow rate – Surge flow rate) / (Design flow rate) Figure 8-a Measured and predicted performance map.595 which is about 6% higher than the intended design value. an excellent agreement can be observed between the measured and the predicted performances across the map. the results of the 1D performance predictions. For comparison.75. The measurements were carried out at five rotational speeds ranging from 44350 rpm to 77100 rpm. This corresponds to an isentropic pressure rise coefficient of 0. the stage achieved a peak total-to-total efficiency of above 80% at a pressure ratio of 2. The variations of the stage total pressure ratio and total-to-total efficiency against the inlet volume flow rate are presented in Figure 8. As no test data in this region of the design space was previously available to the authors. using the method of Casey and Robinson (5). The preliminary design methods used for the design identified many of the difficulties expected from such an unconventional design. stage efficiency versus volume flow. Based on the mass flow rate at peak efficiency condition.J.. Numakura. Casey. R. 63 . ASME Paper GT2014-25378. a surge margin of just above 6. UK. in particular the need for a mixed flow stage. and Casey. H. (2014). 031035 [3] Hazby. 7 REFERENCES [1] Casey. The test data has been used to improve the prediction capability of the Casey Robinson map prediction method for such stages. Figure 8-b Measured and predicted performance diagrams. could also contribute to narrow the flow range at this extreme design condition. 6 CONCLUSIONS The conceptual design and testing of a mixed flow compressor stage in a highly unconventional region of the design space with an extremely high flow coefficient and a high pressure rise coefficient is described. M. Vol.. the application of the compressor for practical applications at such high flow coefficients may be possible only through the application of Map Width Enhancement devices. “A Transonic Mixed Flow Compressor for an Extreme Duty”. The final testing of the stage identified the need for some form of recirculating bleed system to increase the operating range of the stage. Robinson C... The paper shows some of the issues related to the conceptual and preliminary design of the stage. M. “The Design Space Boundaries for High Flow Capacity Centrifugal Compressors” ASME Journal of Turbomachinery. June 14-18. Proceedings of ASME Turbo Expo 2010.V. ASME Paper GT2010-22549 [2] Rusch. Other design features. Tamaki.V..V. D. and Zwyssig. C.. Therefore. such as low backsweep angle and possible flow separation in the curved vaneless diffuser.1% was achieved by the compressor. (2013). the expected narrow operating range and the difficult mechanical issues. M. 135 (3). (2010). 2010.. Glasgow. H. “The Cordier line for mixed flow compressors”. Such a narrow flow range is typical of high speed compressors with highly supersonic flows at the inducer tip section. V... and Robinson. M.. (2010). Düsseldorf.J. ETC10. L. (2010). ASME Journal of turbomachinery.. Butterworth-Heinemann. C.. C. and Hall C.V. 132. 10th European Turbomachinery Conference.J. (2013). “The design of ultra-high speed miniature centrifugal compressors”.A. Finland.. submitted to ASME Turbo Expo 2014: June 16 – 20.V. M. Vol. (2013) “A Method to Estimate the Performance Map of a Centrifugal Compressor Stage. ASME Journal of Turbomachinery. D. and Robinson. Krähenbuhl. 2014.. April 2013 64 . M. “A new streamline curvature throughflow code for radial turbomachinery”. S. C. Germany [4] Dixon. 2010. 135 (2). 6th Edition [5] Casey. Fluid mechanics and thermodynamics of turbomachinery. April. 021034 [7] Casey. Lappeenranta. 2010 [6] Casey.. Boston/Oxford. Zwyssig. 1 INTRODUCTION Today turbocharging has become a fundamental technology to realize engine downsizing. recirculation bypass of the inducer is employed for some applications(2). High efficiency and wide operating range are strongly required for the automotive turbochargers(1). _______________________________________ © The author(s) and/or their employer(s). relatively wide operating range and cost benefits. industrial compressors and turbo shaft gas turbine engines because of their high pressure ratio.Aerodynamic design optimization of a centrifugal compressor impeller based on an artificial neural network and genetic algorithm S Ibaraki 1. I Tomita 1 1 Mitsubishi Heavy Industries. K Sugimoto 1. Especially centrifugal compressors for automotive turbochargers should operate with high efficiency from the surge limit to the choke limit. R Van den Braembussche 2. Belgium ABSTRACT Centrifugal compressors are applied to turbochargers. which makes the understanding of the loss generating mechanisms difficult and requires a considerable design effort to reach good performance. The internal flow in a centrifugal compressor impeller is however three dimensional and shows very complex flow phenomena. In this study a centrifugal compressor impeller for an automotive turbocharger was designed by means of an aerodynamic design optimization system composed of an artificial neural network (ANN) and a genetic algorithm (GA). which is an attractive strategy for low carbon emissions. The design effort can however be reduced by applying an advanced design optimization system as an alternative to a conventional manual design based on the experience of the designer. This resulted in two newly designed centrifugal compressor impellers which were further studied both numerically and experimentally. Ltd. T Verstraete 2. Furthermore two stage turbocharging(5) has been applied to give a considerable wider operating range. Japan 2 von Karman Institute for Fluid Dynamics. Z Alsalihi 2. Especially turbocharger compressors impose a challenge to the designer when both a very wide operating range and high efficiency are required. 2014 65 . Also some devices such as variable inlet guide vanes(3) and low solidity diffusers(4) have been studied. One has higher efficiency with slightly wider operating range compared to the baseline impeller. To increase the operating range of centrifugal compressors. The other one has a twice as wide operating range compared to the baseline impeller with a minor decrease in efficiency. This paper describes the concepts and the procedure of the design optimization system based on an ANN and a GA. the atmospheric pressure is imposed as total inlet pressure. This Base procedure is repeated for a given number of iterations. This resulted in two new centrifugal compressor impellers which were further examined.2 Boundary conditions The compressor is intended to work for atmospheric conditions and should deliver a total to static pressure ratio of 1. all above countermeasures need extra cost and complexity of the turbocharging system. This method requires to specify an optimal blade loading to generate the blade profile.15K. design optimization methods composed by an ANN and a GA requires no prescribed aerodynamic parameters. Therefore.with reasonable stable operating range. which makes it increasingly more difficult to improve the efficiency further. At a result. Therefore. Once the GA has found an Analysis Performance Prediction optimum it is verified by a Navier Stokes solver(11) and is added to the database. In general. centrifugal compressor design can benefit largely from modern optimization techniques. Experimental and numerical investigations of the newly designed impellers are presented and discussed.1 Optimization procedure Figure 1 shows schematically the design Requirement procedure.(13). it is very difficult to specify the optimum blade loading.However. At the outlet a static pressure is imposed corresponding to the required pressure ratio. 2 OPTIMIZATION STRATEGY Start Geometry 2. In this study a centrifugal compressor impeller for automotive turbochargers was designed by means of an aerodynamic design optimization system composed of an ANN and a GA. Inverse design method have been applied to centrifugal compressors and other turbomachinery as optimization methods(10). A more complete description Stop of the optimisation method can be found Figure 1: Optimisation procedure in published papers (12). and the total inlet temperature is fixed at 293. 2. Modern. The loss generating mechanisms in a centrifugal compressor are very complicated and have not been revealed completely. high performance centrifugal compressor impellers demonstrate complex 3D blade geometries with a highly complex 3D flow(6)-(9). The efficiency of centrifugal compressors has been improved over several decades by the progress of computational fluid dynamics and experimental fluid dynamics. The other one has a twice as wide operating range compared to the baseline impeller with a minor decreased efficiency.53. The blade loading is one of the key parameters affecting the compressor performance. One has a higher efficiency with a slightly wider operating range compared to the baseline impeller. 66 . The inlet flow is axially oriented. designers have difficulties to specify the optimal blade loading to achieve a performance improvement. The optimisation is driven by a GA GA in which the performance of each Geometry Blade Generation geometry is analysed by means of an ANN ANN trained on the information contained in a 3D NS database. Alternatively. Performance Data Learning resulting in a more accurate ANN. are given as design constraints. The diffuser meridional contour is defined by second-order Bézier curves and the intermediate Bézier control point B is computed to have a parallel diffuser downstream of that point and smooth trailing edge curve.R1) and 2 (X2. The diffuser exit diameter is fixed equal at the hub and at the shroud. The hub axial position and the shroud leading edge diameter are design variables.Rin). The axial position of point 1 and both coordinates of point 2 are design variables. The parameters β0 and β3 are the metal angles at the leading and trailing edges. The maximum lean angle at the leading edge and trailing edge. The geometry is defined by the meridional contour and the camberline blade angles of the full and splitter blades at hub and shroud. The inlet contour at the hub and shroud is defined by a second-order Bézier curve. The diffuser exit width is also fixed. The trailing edge diameter is set equal at hub and shroud. A and 0 (X0.R2) are design variables. denoted by 0.2. which is a design variable. Figure 2: Meridional contour definition Figure 3: Blade angle definition The blade meridional contour at hub and shroud are defined by third-order Bézier curves with 4 control points. The points denoted by in (Xin. The leading edge axial position of the full blade at the shroud is set equal to the leading edge at hub. (u ) 0 (1 u )3 1u (1 u ) 2 2u 2 (1 u ) 3u 3 (1) Here u is a parameter between 0 at the leading edge and 1 at the trailing edge. β1 and β2 are intermediate parameters in eq. 1 and have no physical meaning. 2 and 3. 1.3 Geometry definition The impeller is backswept with splitter blades. Ule is the percentage of meridional length where the leading edge starts. The leading edge position of the splitter blade is defined by the design variables Ule at hub and shroud. The blade number is allowed to be changed in between minimum 4+4 and maximum 6+6. which are not directly controlled by the used parameterisation. At hub both axial and radial coordinates of points 1 (X1. The trailing edge diameter is not a design variable but it is adjusted to compensate the variation in the blade trailing edge metal angle. The meridional contour definition is schematically shown in Figure 2. The minimum and maximum axial length are fixed.R0) are Bézier control points and constitute the Bézier polygon. where subscript 0 denotes the leading edge of the blade and 3 denotes the trailing edge. The hub inlet and leading edge diameters are fixed. The camberline blade angle definition is achieved using third-order Bézier curves in the form of Bernstein polynomials as shown in Figure 3. These 67 . The Mach number is parametrized at 20 points each. 2. Δmass penalty The difference between the mass flows of each flow channel (divided by the splitter blade) is penalized. The total penalty is the sum of the following items computed using the elements of the performance vector: Mass penalty The difference between the required mass flow and the actual mass flow is penalized. normalized by their limit values (the ANN requires input and output values between 0 and 1). Here wm is the weight of the mass penalty with respecto to the other penalties.0. Initially it was 300 and adjusted during the iterations. The vector of the 27 shape parameters. constitutes the ANN input. 68 . The total performance vector has 5+8*20=165 elements. Here wm2 is the weight of the penalty. The same performance vector is either predicted by the ANN or post-processed from the Navier-Stokes solution.0 (2) Pm wm mref There is a tolerance value of mref/300. max[m req (m1 m 2 )] mref / 300. The ANN penalty is associated to each individual’s performance predicted by the ANN during the optimization cycle. which is the inverse of the total penalty called the objective function. the total-to- total efficiency. Peff weff | tt req | (4) Distortion-skew penalty The diffuser exit flow is desired to be as uniform as possible. The penalties are associated with the Navier-Stokes analysis of the best individual proposed by the GA at the end of each optimization cycle.parameters are the 8 design variables for the full blade camberline at hub and shroud. In both cases the penalty is formulated identically. The description of the blade is completed with the pre-specified thickness distribution.4 Objective functions The GA is based on the evaluation of each individual’s fitness. Any distortion or skew of the radial velocity profile is penalized. along respectively the pressure and suction sides of the blades and splitters. The splitter blades are restricted to have the same trailing edge metal angle as the full blades. 2 m m2 (3) Pm 2 wm 2 1 m1 m 2 Efficiency penalty The difference between the required efficiency (always 100%) and the actual efficiency is penalized. consists of the two mass flows m1 and m2 (see ∆mass penalty below). hence this results in 6 additional design variables for the splitter blades. The efficiency used in the penalty function is the total-to- total adiabatic efficiency. Same pitch at the hub and shroud is imposed at the trailing edge. distortion and skew of the diffuser exit radial velocity profile and full blade and splitter blade Mach numbers.0 with mref=mref for which there is no penalty. The mass flow m1 and m2 are the values for splitter blade pressure side and suction side respectively. Here weff is the weight of the penalty. The performance vector (as well the ANN output) from which the objective is computed. In total 27 design variables define the geometry. 11 is the first derivative of the Mach number and it is performed after the location of the peak Mach number up to approximately the middle of the blade.0) (11) xi xi 1 The summation in eq. 2 ( AB AS ) (8) PL wL AB AS AB and AS are the full blade and splitter blade loading. Mach number deceleration penalty It is desirable to have a smooth deceleration. The loading is calculated as the area between the suction and pressure side isentropic Mach numbers along the normalized blade length. 2 ( M P M min ) (10) Pmach peak wmach peak M P M min The peak Mach number Mp is searched near the leading edge. The velocity at the diffuser exit mid-channel is Vmid. Negative loading penalty It is desired to have a positive loading everywhere. The negative loading is the area where the suction side isentropic Mach number is below the pressure side isentropic Mach numbers. The negative loading penalty penalizes the reversed loading due to higher isentropic Mach numbers at the pressure side of the blade. PNL wNL A (9) Mach number peak penalty It is desired to have a smooth acceleration near the leading edge. The minimum Mach number Mmin is also confined to that interval. 2( M i M i 1 ) Pmach acc wmach acc max( . Pdis skew wdis | d 1 | wskew | s | (5) Here the distortion d and skew s are defined as 2 Vmid (6) d VR VL 2 (V R V L ) (7) s VR VL The radial velocity profile at the diffuser exit is considered. Any reacceleration after the peak is penalized. Mach number acceleration penalty It is desirable to have a smooth deceleration in compressors. Any change in deceleration gradient after the peak is penalized. avoiding high Mach numbers near the trailing edge. 69 . The Mach number peak near the leading edge is penalized. near the hub is VR and near the shroud is VL. Loading penalty The loading difference between the splitter and full blade is penalized. OPT2 has a twice as wide operating range with very suitable characteristic operating curve with a sufficient negative pressure gradient. Figure 5 show the meridional contour of the optimum design impellers with the baseline impeller. 3. at a 1% lower efficiency compared to the baseline impeller. Consequently the authors selected the iteration 17 as the best impeller(OPT1) because it has a good Mach number distribution and equivalent 70 . On the other hand. The maximum blade number is 6+6 in both cases. Table 1 shows the specifications of the optimum design found in both optimization studies compared to the baseline impeller. Table 1: Specifications of optimum design impellers Baseline OPT1 OPT2 OPT1 OPT2 Baseline Figure 4: Optimum design impellers Figure 5: Meridional contour and baseline impeller As described later. A good split of blade loading between the full blade and the splitter blade is also obtained. the Mach number distribution is very smooth. The best impeller is not chosen only for its high efficiency. Moreover. In the second optimization (OPT2) the minimum blade number is 4+4.1 Optimised impeller 1 (OPT1) Figure 6 shows the convergence history and penalty breakdown of the design iterations. Figure 7 shows the comparison of the shroud Mach number distribution of iteration 17 obtained by ANN and NS calculation. OPTIMIZATION RESULTS In this study two optimized impellers were designed by the same optimization procedure. The only difference between both sduties is the blade number limitation. OPT1 has 0. 3. The Mach number distribution. off- design performance and mechanical considerations play as well an important role in the selection process. In the first optimization (OPT1) the minimum blade number is 5+5. A general good agreement is observed. Figure 4 show the optimum design impellers and the baseline impeller. ( M i M i 1 ) ( M i 1 M i ) Pmach dec wmach dec max . 12 is performed downstream the location of the peak Mach number and it is confined to approximately half the blade.5-1.5% higher efficiency with slightly wider operating range compared to the baseline impeller.0 (12) xi 1 xi xi xi 1 The summation in eq. This impeller is one of the best ten designs ordered by the NS total to total efficiency. OPT1 was obtained at iteration 17. iteration 46 was selected as best impeller (OPT2) in this optimization run. almost radial leading edge of the splitter blade.2 Optimised impeller 2 (OPT2) In Figure 8 the convergence history of total to total efficiency is shown.0 0 10 20 30 40 50 60 70 0. The inlet blade height is slightly smaller.2 0. OPT1 has larger inlet blade height and smaller exit blade width compared to the baseline impeller. Iteration 46 is one of the best designs and has the third highest total to total efficiency shown in Figure 8. The blade number of OPT1 is 5+5.4 0. Similar to the selection strategy of OPT1.5 0. Axial length is smaller than that of baseline impeller.8 1.0 Mach number 0.5 NS Efficiency penalty NS Full blade NS Mass penalty NS Splitter blade NS Loading penalty ANN Full blade ANN Splitter blade 100 1.2 0. 150 1.0 Iterations Nondimensional meridional length Figure 8: Efficiency history Figure Figure 9: Shroud 9: Shroud MachMach number number distribution ofdistribution OPT2 71 . Figure 9 shows the shroud Mach number distribution of iteration 46 obtained by ANN and NS calculation. 1. iteration 46 was selected as best because it has not only high efficiency but also a good Mach number distribution and an equal loading split between full blades and splitter blades as shown in Figure 9.5 0 0. The number of blades is 4+4.0 Mach number Penalty 50 0.0 0.6 0.8 1.0 0 10 20 30 40 50 60 70 0.0 Iterations Nodimensional meridional length Figure 6: NS and ANN FigureFigure 7: Shroud Mach number 7: Shroud Mach penalty breakdown distribution numberof OPT1 distribution f OPT1 3.blade loading between the full blade and the splitter blade with a preferable blade geometry for manufacturing.5 NS Full blade 1% NS Splitter blade ANN Full blade Total to total efficiency ANN Splitter blade 1.0 0. A remarkable feature of OPT2 is the forward inclined leading edge of the splitter blade.4 0. Similar to OPT1.6 0. PERFRORMANCE TEST RESULTS AND DISCUSSIONS The optimal shapes resulting from both optimization processes have been further analyzed both experimentally and numerically. On the other hand OPT2 has a very wide operating range with a 1% decrease of compressor efficiency.8 and more than 1. On the other hand. OPT2 has about 1. The operating range is defined as the flow range between maximum flow rate and surge limit at each rotational speed and normalized with the surge limit. representing engine like conditions. OPT1 has a slightly higher efficiency and operating range compared to the baseline impeller. 72 .0% lower efficiency compared to the baseline impeller. In Figure 12 the operating range of the optimum designed impellers and the baseline impeller are compared.5.9 compared to baseline impeller. OPT2 and the baseline impeller.1 Performance test results Figure 10 shows the comparison of compressor characteristics of OPT1.4.0% higher efficiency at pressure ratios above 2. Each compressor volute has the same geometry. The horizontal axis of Figure 12 means the pressure ratio at surge limit at each rotational speed. as well as same exit diameter of the vaneless diffuser.9. Each diffuser has a different width but the ratio of diffuser width and impeller exit width is fixed as same value.5% higher peak efficiency around the total to total pressure ratio of 1. However. Performance tests have been conducted using the compressor inside an automotive turbochager. The operating range of OPT1 is slightly smaller up to pressure ratio 1. 4. The compressor total to total efficiency shown in Figure 10 is normalized by the peak efficiency of the baseline impeller. OPT1 does not have this rapid decrease of its operating range and remains a wide operating range up to a pressure ratio 2. Both optimal impellers are each powered by a turbine driven by heated air or exhaust gases. Compared to the baseline OPT1 has a 0. The compressor characteristics have been measured and compared with the baseline impeller.2. In this study the maximum flow late is defined as the flow rate at which the compressor efficiency drops below 65%. the baseline impeller decreases its operating range rapidly at pressure ratios above 1. Baseline OPT1 OPT2 Figure 10: Comparison of compressor characteristics Figure 11 shows the comparison of the peak efficiency for each rotational speed shown in Figure 10. The flow rate is also normalized by the reference flow rate. A frozen rotor interface between the impeller outlet and the diffuser inlet has been applied. The whole domain of a compressor stage including the impeller.9 but a very wide operating range at pressure ratios beyond 1. The commercial code CFX ver.OPT2 has almost equivalent operating range of the baseline impeller at the pressure ratio below 1.89 (maximum speed: 1.12 was used for this study. OPT2’s operating range has more than doubled at the pressure ratio above 2.2 CFD results The flow phenomena in the optimum designs have been investigated by more detailed steady NS simulation. OPT2 has a very gradual decrease of its operating range with pressure ratio increase. It is obvious that the optimization procedure composed of a GA and an ANN in this study is beneficial to find high performance impellers in a short amount of time. diffuser and volute has been calculated.000 for the diffuser and 250. OPT2 has a suitable compressor characteristic for turbocharger applications which require a wide and stable operating range. stream lines (only in right figures of Figure 13) and the entropy distribution and are compared. 410.000 for the impeller. There is a possibility to get the variety of geometry and characteristic of optimum design such as OPT1 and OPT2 with same design procedure. The difference of the compressor characteristics of OPT1 and OPT2 shows the diversity and complexity of the aerodynamic design of a centrifugal compressor impeller.470. FigureFigure 11: Comparison 11: Comparison of of total Figure 12: Comparison of total to total to total efficiency operating range ffi i Even though OPT1 and OPT2 have been selected as best impellers which show higher efficiency compared to the baseline impeller during the optimization procedure. OPT1 and OPT2 have completely different characteristics as demonstrated by the performance test. 4.0) and the flow rate near the peak efficiency condition. CFD was conducted at the normalized rotational speed of 0. Normalized helicity is defined as cosine of the angle made between the vortex vector and the velocity vector. Also shown in Figure 13 are the identified vortex cores which are coloured with normalized helicity. In Figure 13 the limiting stream lines on the blade surface. and the domain where its absolute value is 1 indicates a strong rolling-up of a streamwise vortex.000 for the volute. The k- ε model was used as the turbulence model.9. OPT1 has achieved higher efficiency compare to the baseline impeller as expected.2 compared to the baseline impeller. On the other hand OPT2 has lower efficiency but has a significant increase of its operating range. Regarding 73 . The number of grid cells is about 2. The discrepancy of OPT2’s efficiency between the estimation and the test result remains to be solved. As a result. Low momentum fluids Low momentum fluids by LE tip leakage vortex by secondary flow and Accumulation of low momentum fluids and tip leakage flow tip leakage flow Hn [-] by LE tip leakage vortex Tip leakage flow Tip leakage flow Secondary flow LE tip leakage vortex Secondary flow Accumulation of low momentum fluids by secondary flow (a) Baseline Hn [-] Tip leakage flow Tip leakage flow LE tip leakage vortex (b) OPT1 Hn [-] Tip leakage flow Tip leakage flow Secondary flow LE tip leakage vortex Accumulation of low momentum fluids by secondary flow (c) OPT2 Figure 13: Comparison of limiting stream lines. it is clear that the baseline impeller has a strong secondary flow rolling up from hub to shroud on the suction surface at the inducer. This low momentum fluids combined with the tip leakage flow more downstream the full blade accumulate and compose the high entropy region near the tip corner of the splitter blade pressure 74 .the entropy distribution the regions which have a relatively high entropy are mapped in streamwise sections in Figure 13. Because of this strong secondary flow. low momentum fluids start to accumulate on the suction surface of the full blade near the end of the inducer. vortex structure and entropy distribution From Figure 13(a). It means that the design optimization system has succeeded to improve the internal flow similar to the design of OPT1. OPT2 has a relatively larger area of low momentum fluid at the exit of the passage of the splitter blade pressure surface and has lower efficiency compared to OPT1. Because this stronger LE tip leakage vortex travels into the passage of the splitter blade pressure surface. As demonstrated above the strong secondary flow motion and tip leakage flow are the main causes of loss generation in the baseline impeller. In contrast to the baseline impeller OPT1 does not have a remarkable secondary flow at the inducer as shown in Figure 13(b). OPT1 has a very smooth flow without pronounced secondary flow at the inducer and thus achieves a higher efficiency compared to the baseline impeller. the accumulation of low momentum fluid is suppressed near the tip corner of the splitter blade pressure surface at the impeller exit compared to the baseline impeller. The LE tip leakage vortex is remarkable different between the baseline impeller and the OPT1 impeller. the accumulation of the low momentum fluid on the full blade suction surface near the end of the inducer is observed. This is due to a wider blade pitch distance as a result of the smaller blade number. the maximum loss of this passage is higher than that of the baseline impeller. This discrepancy between CFD and test results remains to be solved. At the tip corner of the full blade pressure surface at the impeller exit mostly tip leakage flow downstream the inducer accumulates and highest loss region is composed. it is obvious that the high loss region is suppressed clearly compared to the baseline impeller because the LE tip leakage vortex of the full blade leading edge does not come in and the related loss is diminished.surface at the impeller exit. the secondary flow and the accumulation of the low momentum fluid are not as severe. Because of this secondary flow. The internal flow phenomena and loss generation mechanisms of OPT2 are almost identical to those of OPT1. On the other hand. In contrast to OPT1. As shown in Figure 13(a) the LE(leading edge) tip leakage vortex is generated at just downstream the full blade leading edge. According to the CFD result. The investigations to find the mechanisms of map width enhancement of OPT2 are beyond the scope of this paper. However the area of highest loss region is much smaller than that of the baseline impeller because there is almost no accumulation of low momentum fluid by the secondary flow. Owing to this. As a result. This LE tip leakage vortex travels further downstream while accumulating low momentum fluid and composes the high loss region near the shroud of the splitter blade suction surface at the impeller exit. as it only develops more downstream the main blade and postpones its movement into the passage of splitter blade pressure surface. OPT2 has a secondary flow rolling up from mid-span to shroud on the suction surface at the inducer (similar to the baseline) as shown in Figure 13(c). OPT2 has higher efficiency compared to the baseline impeller in spite of lower efficiency confirmed by the performance test. This LE tip leakage vortex moves into the passage of the splitter blade suction surface and composes a low flow region near the shroud. even though LE tip leakage vortex is much stronger due to the increased blade loading resulting from the smaller blade count. 75 . Regarding the passage of the splitter blade suction surface. The authors have found out the breakdown of the tip leakage vortex plays an important role in stabilizing the unsteadiness of the flow near the stall condition in a separate study(14). compared to the baseline impeller. Significant increase of the operating range has been confirmed by the performance test of OPT2. It seems that the spanwise and streamwise blade loading distribution between the full blade and splitter blade are optimized to succeed suppressing the secondary flow by the aerodynamic design optimization system composed of an ANN and a GA in this study. (3) Tomita. and Yamada. K. Samata. S. This successful result clearly demonstrates the benefits of advanced optimization systems composed of an ANN and a GA. The operating range above the pressure ratio 1. Hayashi.9 is wider than that of the baseline impeller. 23rd CIMAC.. Kuma. (5) An. “Development of Variable Two-stage Turbocharger for Passenger Car Diesel Engines”.. B. H. T. (2) OPT1 has a 0. “A New Operationg Range Enhancement Device Combined with a Casing Treatment and Inlet Guide Vanes for Centrifugal Compressor”.2 with almost equivalent operating range at low pressure ratio and 1% lower efficiency compared to the baseline impeller. The internal flow phenomena and loss generation mechanisms of OPT2 are almost identical to those of OPT1. T. T.. Vol. 43. S. The following conclusions are obtained.. Owing to a slightly stronger secondary flow compared to OPT1..8 and more than 1. (7) Ibaraki.2 compared to the baseline impeller. 11th International Conference on turbochargers and turbocharging. Ogita. 1988. Matsuo.. (8) Ibaraki. H. and loading split between the full blade and splitter blade are optimized to suppress the secondary flow and achieve a higher efficiency. “Flow Investigation of a Transonic Centrifugal Compressor for Turbocharger”. CIMAC No. Higashimori. (4) Ibaraki. and Matsuo. 2007. H.0% higher efficiency above pressure ratio 2. T. were designed by an advanced optimization system. 2010. The authors have applied this optimized design systems to the aerodynamic design of centrifugal compressor impellers for automotive turbochargers. (2) Fisher. Jinnai. And Nanbu. “Development of a Wide Operating Range Turbocharger Compressor with a Low Solidity Vaned Diffuser”. K and Suita.. 4. S. OPT2 has a twice as wide operating range compared to the baseline impeller with a minor decrease in efficiency. Vol. OPT1 has a higher efficiency with slightly wider operating range compared to the baseline impeller.5.. CONCLUSIONS In this study an advanced optimized design system composed of an ANN and a GA alternative to a conventional design system has been proposed. H.. “Aerodynamics of a Transonic Centrifugal Compressor Impeller”. “Flow Investigation of a Centrifugal Compressor for Automotive Turbochargers”. 1998. (6) Ibaraki. MHI Technical Review. “Application of Map Width Enhancement Devices to Turbocharger Compressor Stages”. ASME Journal of Turbomachinery. An. OPT2 has lower efficiency... IMechE. REFERENCES (1) Osako. T. SAE Paper 880794. Significant increase of the operating range was investigated in a separate study.. (1) Two optimum design impellers. “Development of the High Performance and High Reliability VG Turbocharger for Automotive Applications”. (3) OPT2’s operating range has increased more than double at the pressure ratio above 2.. S.5% higher peak efficiency at pressure ratio 1.. SAE Paper 98-P94. B. Two impellers have been designed and their performance has been further studied experimentally and numerically. 166. N. 2003. Sumida. F. 76 . Shiraishi. Y. Higashimori. Suzuki. 6. Ibaraki.125. 3. There is almost no secondary flow on the suction surface of the full blade... 2006. Vol. No.. B. and Mikogami. 2001. S. T.. No. OPT1 and OPT2. I.. The breakdown of the tip leakage vortex plays an important role in this. It seems the spanwise and streamwise blade loading. 346-351. A. 47.2，pp. No. H. 2014.. T. and shows that innovative aerodynamic design can be found while speeding up the design time. MHI Technical Review. “Rotor-Stator Interaction Analysis Using the Navier-Stokes Equations and a Multigrid Method”. I. (14) Tomita. 326-332).. 1995. and Van den Braembussche R. Istanbul. (11) Arnone A. K. 1998. S. ASME Paper No. M. and Van den Braembussche R.. Iwakiri.. 2002. Furukawa. 2012. “The Effect of Tip Leakage Vortex for Operating Range Enhancement of Centrifugal Compressor”. V. and Takahashi. M. ASME Paper No. GT2012-68947. M. C. “Vortical Flow Structure and Loss Generation Process in a Transonic Centrifugal Compressor Impeller”. “Evaluation of a Design Method for Radial Impellers Based on Artificial Neural Network and Genetic Algorithm”.. (10) Zangeneh. Roduner. S. K.. (13) Alsalihi Z. 121.. (pp. ASME Trans. Journal of Turbomachinery.(9) Ibaraki. (12) Pierret S. Yamada. Furukawa..A.A. “Improving a Vaned Diffuser for a Given Centrifugal Impeller”. 9... ASME Paper 95-GT-177. ASME ESDA 2002/ATF-069.. ASME Paper. GT-2002-30621. 2002. GT2007-27791. No. and Pacciani R. D. K. Vol.. Ibaraki. 77 . 2001. “Turbomachinery blade design using a Navier-Stokes solver and Artificial Neural Network”.. these regions greatly affect the expected performance of the compressor. J Early School of Mechanical & Aerospace Engineering.Inlet recirculation in automotive turbocharger centrifugal compressors P X L Harley. The paper improves upon an existing correlation between the rate of development of the recirculating region and the compressor stage. the point at which the recirculating flow begins to develop and the rate at which it grows are investigated. 2014 89 . This study analyses the inlet recirculation region numerically using several modern automotive turbocharger centrifugal compressors. Compressors with relatively large map widths tend to have very large recirculating regions at the inlet when operating close to surge. which is supported by results from the numerical analysis. Germany ABSTRACT As the designers of modern automotive turbochargers strive to increase map width and lower the mass flow rate at which compressor surge occurs. Using 3D Computational Fluid Dynamics (CFD) and a single passage model. UK D Filsinger. All numerical modelling has been validated using measurements taken from hot gas stand tests for all compressor stages. the recirculating flows at the impeller inlet are becoming a much more relevant aerodynamic feature. Queen’s University Belfast. NOMENCLATURE ܣ Area Subscripts ܤ Blockage (-) ܾ Blade ܥ Absolute velocity (m/s) crit Critical ܦ Diameter (m) ݉ Meridional DF Disk Friction max Maximum ݉ሶ Mass flow rate (kg/s) 0 Stage inlet OP Operating Point 1 Impeller inlet PR Pressure Ratio (-) ߠ Tangential ܷ Blade velocity (m/s) ܸሶ Volumetric flow rate (m3/2) VL Volute Loss ߚ Relative angle to meridional (deg) ߟ Efficiency Total-Total (-) ߩ Density (kg/m3) ߶ଵ Stage flow coefficient (-) _______________________________________ © The author(s) and/or their employer(s). M Dietrich IHI Charging Systems International GmbH. S W T Spence. however the spin up rate of the compressor can be limited by the surge margin of the stage. most notably in the pressure ratio map.1. also known as ‘trim’. Inlet recirculation blockage in this case is defined as the percentage of the inlet area that does not contribute to the stage mass flow rate as shown by Figure 2. although the drawback is significant recirculation at the inlet to the impeller driven by adverse pressure gradients along the shroud. Typical centrifugal compressor performance dictates that the surge mass flow rate increases with rotational speed. during the spin up the compressor must not operate below the surge mass flow rate. Modern automotive turbocharger centrifugal compressors are designed to provide boost as quickly as possible when power is demanded to improve driveability at low engine speeds. Some ways of achieving this are by applying backsweep to the impeller trailing edge and increasing the inlet-to- outlet radius ratio. The blockage caused by the recirculating flow continues to grow in size toward the surge region of the map. The recirculating flow causes an aerodynamic blockage which affects the incoming flow to the compressor. The result is an improved surge margin. It would be ideal to have instant boost and hence instant torque. Reducing the inertia of the rotating components is a continual design target. INTRODUCTION A historic problem with turbocharged passenger cars was turbo lag. New design methods have reduced the mass flow rate at which compressor surge occurs while maintaining the flow range from surge to choke. Figure 1 CFD compressor map C-1 showing contours of impeller inlet blockage (%) as a result in recirculation 90 . but due to inertial effects this is not possible. inlet recirculation is addressed in this paper at steady state operating conditions. Recirculation has a very notable effect on the compressor stage performance. The effect on the pressure ratio map is evident in Figure 1 where a deviation from the trend line is seen at the mass flow rate that inlet blockage begins. thus avoiding the unstable surge region of the map. Convergence was defined when the RMS residuals fell below 1e-4.01%. MODELLING Due to the difficulty in measuring the size of the recirculating region via experiment. and the rotating domain was meshed using Turbogrid and contained approximately 1. Figure 2 Typical meridional streamlines showing impeller inlet recirculation Typically the recirculating body of fluid fills a portion of the meridional passage similar to that shown in Figure 2. The inlet and vaneless diffuser domains were stationary and impeller domain was rotating. the majority of the passage had a y+ of less than 2 at all operating conditions. henceforth referred to as C-1. The geometry and mesh used in the simulation of the three stages have been developed so as to ensure comparability. Total cell count for all three stages was approximately 1. This study investigates the impeller inlet recirculation for three automotive turbocharger centrifugal compressors. and momentum in all domains fell below 0. C-2 and C-3. The inlet domain was made long enough so as to ensure that even at the most extreme operating condition the inlet recirculation did not cross the inlet boundary. The stationary inlet and vaneless diffuser domains were meshed in ANSYS ICEM and had approximately 200k and 100k cells respectively.05%. The recirculation zone is almost self-sustaining with regard to mass flow. A single blade passage was simulated using the ANSYS CFX RANS solver and the geometry prepared using the dimensions shown in Figure 3. The stage mass flow is very similar to the mass flow rate passing through the unblocked active flow entering the impeller. Frozen rotor interfaces were used between the stationary and rotating domains. energy. the imbalances of mass.2million cells. 91 . CFD (Computational Fluid Dynamics) was used to analyse the flow feature numerically. The CFX SST (Shear Stress Transport) turbulence model was used which required a maximum y+ (dimensionless wall distance) of less than 5. 2.5million which had been reached through a mesh independence study. normally within a couple of percent based on recent numerical studies. and the total- total isentropic efficiency was fluctuating less than ±0. The flow feature is analysed qualitatively and shows how the size of the inlet recirculation varies predictably with respect to stage geometry. 1 1D loss modelling In order to ensure comparability of the single passage CFD results with test data (presented later in Section 2. 92 . All compressor maps have been non-dimensionalised using the maximum pressure ratio and mass flow rate for each complete compressor data set of CFD and test results. Some error exists in the simulated pressure ratio for C-2 (Figure 5). All hot gas stand data was collected in accordance with SAE J1826 [4]. The main differences between the CFD and the test data is the presence of a scroll/collector. this is potentially related to the interaction of the simulated stage with the scroll volute.5D R1 S1 1mm D z -ve 0 Figure 3 CFD compressor single passage model dimensions 2.2). To compensate for this. The impeller back disk creates friction with the housing adjacent to it due to viscous shear of fluid in a small space. 2.2 CFD validation The CFD results with and without the 1D loss models applied are plotted in Figures 4. S2 4D 0. The meanline volute modelling method of Weber & Koronowski [1] was used to model the effect a volute would have on stage performance. and 6 against hot gas stand data. Another difference between the single passage model and the test data was the lack of an impeller back disk in the CFD model. certain losses had to be applied to the CFD results. the disk friction (DF) loss of Aungier [2] was used as it is the most representative of the classic Daily and Nece [3] study. Surge is defined on the test rig using pressure fluctuations in the compressor discharge line to ensure repeatability. Good pressure ratio and efficiency correlations between the test data and CFD simulations are shown for C-1 and C-3 in Figures 4 and 6 respectively. 5. In the turbocharger compressors used in this study a scroll volute is the norm and for that reason a volute model had to be applied. 9 0.4 0.6 0.9 1.7 0.0 1. whereas C-1 and C-2 have a similar drop in peak efficiency at the lowest tip speed of almost 20 points.9 CFD without VL & DF CFD with VL & DF 0.8 0.0 1.6 0.3 Data0.6 0.0 0. 1.0 0.4 0.8 η/ηmax [-] 0. As expected the largest of the three compressors.4 0.1 0.2 0.0 0.6 0.1 0.7 0. The reduction in efficiency is the result of the inescapable heat transfer inherent with hot gas stand testing as heat is exchanged between the hot turbine and relatively cool compressor through the bearing housing. C-3 has the smallest efficiency decrement of only approximately 7 non-dimensionalised percentage points. and 6).1 0.0 20% 0. 93 . 5. The same effect is present in all three compressor maps (Figures 4.8 0.5 1.8 43% 0.7 0.5 0.4 0.7 PR/PR max [-] 0.5 0.2 Test C-1 0.5 0.2 0.9 1.3 0.1 ṁ/ṁmax [-] Figure 4 C-1 compressor map comparison The test data shows a drop in efficiency at low tip speeds which is not supported by the CFD.3 0. In the C-1 and C-3 simulations the position of the typical ‘kink’ in the surge margin is captured very well.0 1. this time a maximum of ~65%.9 1.6 0.2 0.5 1. For C-1.1 CFD without VL & DF 1.5 0.6 0.5 0.3 0. 94 .0 1.8 0.5 0.0 0.7 C-20.0 CFD with VL & DF 0.1 0.4 0.7 0. The surge mass flow rate CFD prediction for C-3 in Figure 6 is much better with a maximum error of ~16%.0 0.9 Data 1. 1.1 ṁ/ṁmax [-] Figure 5 C-2 compressor map comparison The surge margin in the CFD was defined as the mass flow rate that the solver would no longer converge to the criterion described previously.6 0. The maximum error in the CFD surge mass flow rate prediction for C-1 is ~43% as shown in Figure 4.1 0.7 0.6 0.4 0.7 0.8 0. surge at low and high tip speeds was predicted very well but the error increases from low to mid tip speeds.4 0.9 0.2 0.4 0. For C-2 in Figure 5 considerably more error is present in the surge margin prediction. again this is potentially related to the simulation not containing the asymmetric volute found on the tested compressor.9 PR/PR max [-] 65% 0.8 Test 0.1 0.3 0.8 η/ηmax [-] 0.0 20% 0. All Turbo Line data was mass flow averaged circumferentially.5 0.3 CFD post-processing The ‘Turbo Mode’ within ANSYS CFX Post was used to develop meridional plots which were used to analyse the growth of the recirculation flow paths using streamwise velocity plots.6 0.1 ṁ/ṁmax [-] Figure 6 C-3 compressor map comparison 2.1 C-3 0.2 0.4 0.3 0.1 CFD without VL & DF 0.8 16% OP6 PR/PR max [-] 0.7 0. The sample points were then used to calculate an average flow property for the area contained between the sample points and a mass flow rate calculated. 1.7 OP5 OP4 OP3 OP1 OP2 0.6 0.9 7% 0.8 η/ηmax [-] 0.1 0.4 0.9 1.8 0.0 0.5 1. The sample points were distributed so as to allow equal area sample regions. The recirculating limit shown in Figure 7 defines the size of the inlet blockage. the mass flow rate through each circular segment was cumulatively summed until zero mass flow was achieved.5 0. the ANSYS ‘Turbo Line’ feature was used to extract relevant data very close to the leading edge of the impeller.7 0.0 1.5 0. 95 . To find the extent to which the blockage fills the impeller inlet the axial velocity and density was extracted at 200 sample points from hub to shroud.8 0.4 0.4 0. The mass flow rate recirculating close to the shroud is considered to be negative. Starting from the shroud. Also.6 0.6 0.0 0.3 0.0 0.0 0.9 1.0 1.9 CFD with VL & DF 0.2 Test Data0.7 0.3 0. 17 Out Cm1 [m/s] (bottom axis) Recirculation Limit Density [kg/m^3] (top axis) Recirculation span [-] In Active Flow -40 -20 0 20 40 Figure 7 Typical low tip speed low mass flow rate impeller inlet flow conditions The velocity profile shown in Figure 7 is similar to that proposed by Qiu et al. but it is encouraged by the 96 . Figure 8 shows meridional plots of streamwise velocity at six operating points (OP) with the contours limited to show positive (white) and negative (black). The variation of spanwise density in the recirculation zone shows an increase toward the shroud driven by the radial pressure gradient of the swirling flow.15 1. The six operating points in question are plotted in Figure 6 for reference.1 Recirculation initiation In order to demonstrate the growth of the recirculation zone a series of meridional plots are shown for compressor C-3. Based on the knowledge developed the Qiu et al.14 1. The density toward the hub is maintained by the incoming active flow. and that using the Turbo Line method as described will miss the onset of recirculation due to the current definition (Figure 7). RESULTS AND DISCUSSION As mentioned earlier the ‘Turbo Line’ used was close to the leading edge of the impellers as represented in Figure 7. 1. It was realised that the recirculation actually starts within the impeller. [5] and is therefore supported by the CFD simulations. [5] method is improved upon to provide a more accurate estimation of blockage for an automotive turbocharger style compressor. The recirculation seen in OP2 (operating point 2) is driven by the adverse meridional pressure gradient at the shroud. 3. 3.13 1. the recirculation starts along the shroud surface within the blade passage. To ensure that the method was valid a series of meridional plots were used to judge the starting position of the recirculation and compare this to the blockage seen at the Turbo Line. As is clear from Figure 8.16 1. 3. hence the recirculation fills a portion of the passage and maintains the ratio of 1. It is based on the ‘Two Element In Series’ model of Japikse [6]. the inlet stage flow coefficient was used as given by Equation 2. with further reductions in mass flow rate the recirculation zone grows.5. although the same process could be used to measure and quantify the outlet recirculation and blockage. Qiu et al. This recirculation zone is not analysed here. [5] produced a blockage correlation based on a critical area ratio. In order to compare the critical area ratio for all operating conditions and all compressors. At OP2 the recirculation zone does not actually cross the Turbo Line and therefore the post processing method outlined earlier will not register this as an operating condition with blockage.5 in the active flow region. It assumes that the maximum diffusion area ratio that can be maintained stably is 1. and hence the blockage grows. 97 . The recirculating zone extends not only in the spanwise direction. The dotted line in Figure 8 represents the position of the Turbo Line placed in ANSYS CFX.tip leakage flows. [5] there is also recirculation at the trailing edge of the impeller. As expected. Blockage is defined in this model by Equation 1 [5].2 Blockage correlation In a recent paper. =1− − 1 [1] ( − ) cos ( ) where cos = cos In order to analyse this model the CFD results were used to calculate the critical area ratio required to provide the inlet blockage seen in the results. OP1 OP2 OP3 OP4 OP5 OP6 Figure 8 Meridional plots showing regions of negative streamwise flow (see Figure 6 for corresponding operating points) As is outlined by Qiu et al. but also in the axial direction (although this is not investigated here) allowing the Turbo Line to detect the recirculation. 10 0. 2. = 160 − 25 + 2. the improved Qiu et al. With a varying critical area ratio.02 0. [5] model with a varying critical area ratio.4 2. = [2] The results in Figure 9 show that a constant value of critical area ratio is not supported by the CFD simulations. although the target here was a model that worked well with automotive turbocharger style centrifugal compressors. All compressors appear to fall on to this common trend line.08 0.00 0.04 0. The recirculation starts at an area ratio of approximately 1.8 1.0 0.6 1. The correlation can then be used to compare the CFD blockage against the blockage now predicted using the Qiu et al. The original Qiu et al. [5] model by correlating the critical area ratio against the flow coefficient.3 as opposed to the 1.5 value Qiu et al. [5] model is shown in their paper to have good general agreement with a range of impeller types.06 0. [5] suggests. [5] model is shown in Figure 10 to have very good agreement with the CFD simulation. The data in Figure 9 was used to further develop the Qiu et al.12 0. [5] inlet recirculation method 98 . Instead a trend is shown to increase in a parabolic shape toward lower flow coefficients.2 1. the result is Equation 3.4 1.14 ϕ01 [-] Figure 9 Critical area ratio required to provide blockage from CFD using Qiu et al.0 Qiu Current ARcrit [-] 1.2 C-1 C-2 C-3 2.2 [3] >0 The current correlation is plotted on Figure 9 and is only valid when the calculated inlet blockage is greater than zero. The recirculating flow region is almost self sustaining with regard to mass flow rate. [5].15 0.05 0.3 0. an active flow region and a recirculating zone. The results are compared with a recent blockage prediction model by Qiu et al.4 C-2 (Current) C-3 C-3 (Qiu) B1 [-] C-3 (Current) 0. the CFD does not support this model.10 0. For this reason it could be assumed that the flow passing through the active flow region wholly contributes to the flow conditions at the impeller outlet.20 0.0 0. [5] model and current 4. CFD results are used to show that this coincides with impeller inlet recirculation which in turn causes an aerodynamic blockage.25 ϕ01 [-] Figure 10 Comparison of CFD inlet blockage with Qiu et al. CFD simulations are carried out for three automotive turbocharger centrifugal compressors and the data used to analyse the rate at which the inlet blockage grows. CONCLUSIONS Recirculation is a flow feature that has a significant impact on compressor performance.2 0. 0. The result is two distinct regions of flow. 99 . The weakness of the Qiu model was identified to be the assumption of a constant critical area ratio.00 0.1 0.5 C-1 (Qiu) C-1 (Current) C-2 C-2 (Qiu) 0. Pressure ratio trends in compressor test data can show a deviation from the expected trend when operating at relatively lower mass flow rates.6 C-1 0. Recirculation can exist at the inlet and exit of the impeller although this study focuses on the inlet recirculation. and the mass flow rate passing through the active flow region is very similar to the stage flow rate. USA. 86-GT-216. ASME Conference Proceedings. Vermont. “Centrifugal Compressors: A Strategy for Aerodynamic Design and Analysis”. ASME. D. Volume 82. ISBN 0-7918-0093-8..R. Concepts NREC. Thanks are also given to ANSYS Inc for the use of their software for numerical modelling in this research programme. USA. J..H. Japikse. Journal of Basic Engineering. Proceedings of ASME Turbo Expo 2008: Power for Land. 1986. REFERENCES [1] Weber.. The new correlation is shown to be supported by the CFD results. 1996. [5] method was still used but with a varying critical area ratio. M.W.. Pennsylvania. New York. [2] Aungier. 100 . [5] Qiu. & Anderson.. 1960. GT2008-51349. ACKNOWLEDGMENTS The authors would like to thank IHI Charging Systems International for their technical support and provision of the required compressor geometry and test data.E. Germany. “Centrifugal Compressor Design and Performance”.. R. M. New York. [4] SAE J1826 March 1995 “Turbocharger Gas Stand Test Code”. [3] Daily.E. The Qiu et al. “Chamber dimension effects on induced flow and frictional resistance of enclosed rotating disks”. D. X. “Meanline Performance Prediction of Volutes in Centrifugal Compressors”. ISBN 0-933283-03-2. [6] Japikse. USA. USA. New York. ASME Press.. Berlin. & Koronowski. R. 2008. 2000. Sea and Air..The critical area ratio was analysed and correlated against a flow coefficient. C. USA. “A Meanline Model for Impeller Flow Recirculation”. & Nece.. The flow process is therefore split into one or two heat transfers at constant pressure and an adiabatic compression/expansion. but has a significant influence on the outlet temperatures of the device. 8) use the _______________________________________ © The author(s) and/or their employer(s). On the turbine side heat transfer can represent 20-90% of the total enthalpy drop. This study shows the importance of undertaking a full thermal characterisation and the need for accurate adiabatic maps in turbocharger simulations. UK 2 CMT – Motores Térmicos. On the gas stand. 1 INTRODUCTION The accuracy of turbocharger models needs to increase to allow more use of simulation tools in engine development. 2014 103 . The major findings from these investigations are that the heat transfer has a limited effect on the flow behaviour of the turbocharger turbine and compressor. Most of this heat is then lost to ambient with the remainder transferred along the bearing housing. as temperature are used to measure work. Universitat Politècnica de València. derived from steady state measurements taken on steady flow facilities. F J Arnau2. Typically heat transfer is ignored in turbocharger models and these are map based. When these maps are used on the engine this leads to poor estimation of operating speed. Shaaban (3) and Romagnoli and Martinez-Botas (5) provide an analytical solution to heat transfer in the bearing housing simplifying the geometry to a series of cylinders and fit results for an individual device for other model unknowns. P Olmeda2. Spain ABSTRACT A lumped capacity heat transfer model has been developed and compared to measurements from a turbocharger operating on a 2. University of Bath. The model parameters have been estimated based on similar devices and this study quantifies the errors associated with this approach. gas temperatures and air flows. M Reyes-Belmonte2 1 Department of Mechanical Engineering. the heat transfer is more complex as unless the turbocharger is insulated. this gives poor estimates of the aerodynamic performance. A sensitivity study showed the parameters of the heat transfer model influence gas temperatures by only ±4oC but housing temperatures by up to 80oC. Other authors (7. Turbine outlet gas temperature prediction was improved with RMSE reduced from 29.Modelling of turbocharger heat transfer under stationary and transient engine operating conditions R D Burke1. heat transfer from the turbine to the compressor is small (6). primarily to oil and cooling water if present.2L Diesel engine under steady and transient conditions ranging from 1000-3000rpm and 2-17bar BMEP.5oC to 13oC. Various models have been proposed to account for heat transfer each employing the assumption that work and heat transfer occur independently in the compressor and turbine. A number of studies into the effect of heat transfer have been published over the past 10 years (1-5). Transient simulations showed how errors in the thermal capacitance also lead to errors. On the compressor side. Working fluids in turbocharger model have been named from here on as Gas (exhaust gases moving the turbine).lumped capacitance method. C (compressor case) and three nodes for the turbocharger bearing housing named H1 (placed near turbine case). This heat can flow to ambient or along the bearing housing towards the cooling oil and further ambient losses. In both cases. It includes five metal nodes that are representative for the geometry of the turbocharger. conduction and radiation. this heat flow can reverse. These are T (turbine case). the difficulty lies in determining the thermal resistances between the different nodes. In addition to work. 2 MODEL DESCRIPTION 2. If correctly parameterised. these models offer the potential to reduce exhaust gas temperature prediction errors from +40oC to ±10oC (6). That evidence allows the simplification of heat transfer problem inside the turbocharger considering it as a one-dimensional problem (9) instead of the three- dimensional case. Serrano et al (9) propose a method to calculate conductivities and capacitances using the turbocharger housing as a 1D heat flux probe.1 Model Overview An overview of energy flows in turbochargers is shown in figure 1 (a). under high compression ratios. Oil (lubricating oil) and Ambient. heat is also transferred between turbine and compressor. Depending on the operating conditions. In this approach. heat may continue to be conducted to the compressor housing or. In order to simplify the complex heat transfer 104 . in this work the model is applied to a turbocharger operating on an engine system where detailed model parameters were not available. non-adiabatic/externally insulated and non-adiabatic/non- insulated maps to derive the internal convection and external heat losses. splitting the turbocharger housing into a number of thermal nodes linked through convection. Air (compressed air by the compressor). Figure 1 (b) shows the proposed 1-D lumped model to account for heat transfer effects inside an automotive turbocharger. significant experimental effort on dedicated test facilities is required to parameterise the models. Turbine Compressor Wf W Wc T QG/T QT/B QC/A QB/C Q B/Oil QB/Amb QC/Amb QT/Amb (a) (b) Figure 1: (a) Work and heat flows in the turbochargers and (b) lumped capacity thermal model Several studies have demonstrated that radial temperature distribution in a cross sectional area is negligible compared to the axial temperature distribution (10). However. That division is justified by the high temperature gradient across the housing. H2 (placed in the central part. where oil and water comes to the turbocharger) and H3 (placed near compressor case). They then compare adiabatic. Compressor node (C) is connected to ambient (QC/Amb). Housing node H3 in connected by conduction to node C (QH3/C). Conductive conductances are represented with thermal resistors drawn in black in figure 1 (b). the temperature after adiabatic compression can exceed the temperature of node C and this heat flow reverses. QH2/AMB and radiation). to ambient (QH3/AMB) and nodes T and C by radiation. QH2/Oil. Metal nodes are connected by means of conductive conductances (Ki. the temperature after adiabatic compression (node Air) is lower than node C (low loads): in this case the compressed air received heat (QC/Air). H1. It is worth noting that for some operative conditions. Q icond . H2 and H3. An additional term for radiation heat transfer is included for the heat transfer between nodes and ambient taking into account the view factors between nodes from simplified geometries (11) (light grey resistors in figure 1 (b)). and by forced convection with a node placed at compressor diffuser (QC/Air). From node T.i) and heat transfer is calculated using Newton's cooling law (equation 2). It is assumed that heat is first exchanged by forced convection between nodes H1 and oil (QH1/Oil). Heat flows in the model are summarised by the following process: A proportion of exhaust gases energy is transmitted to the turbine case (node T) before the turbine stator and rotor. Connection between metal and fluid nodes is represented by bold grey thermal resistors in figure 1 (b). the oil temperature rises due to frictional losses (Wm). j ·Ti T j 1 Metal nodes are connected to fluid nodes by means of convective conductances (hAl. heat flow can also reverse in the bearing housing. At node H1. to ambient (QH1/AMB). 105 . That phenomenon has been taken into account in the network model by means of capacitors (represented in light grey). forced convection to oil (QH1/Oil). This energy (QGas/T) reduces the temperature governing the expansion process. further heat is transferred by forced convection with node H2 (QH2/Oil). radiation with nodes T.j Ki . or radiation with nodes T and C. Depending on the relative temperatures and thermal resistances connecting nodes C.i · Tl Ti 2 Metal nodes can store energy during transient processes as their temperature increases or decreases due to their associated mass. The opposite is true on compressor side and it has been assumed that heat transfer phenomena occurs after the compression work. Node H2 is similar to node H1 for non-water-cooled devices and similar heat flow paths are defined (QH2/H3.j) and the heat flow between two nodes is calculated using Fourier's Law (equation 1). Conductive conductances between adjacent metal nodes in the turbocharger are constant for any operative condition since that property depends only on the internal geometry and cases material.phenomena in the turbine side and taking into account the exposed areas. H2. Q lconv . At higher loads. or via radiation to nodes H1. heat flux can be transmitted following these possible paths: conduction to node H2 (QH1/H2). secondly. The oil temperature rises both as a result of heat transfer and frictional losses. Finally. heat is conducted to node H1 (QT/H1) or transferred to ambient (QT/AMB). H3 and Ambient. H3 and C. it has been assumed that the heat transfer phenomena occurs before the expansion process as it is shown in the left h/s diagram of figure 1 (b).i hAl . The 106 .5mm k-type thermocouples and turbocharger instrumentation of speed was measured using an eddy current blade turbocharger count device from Micro-Epsillon. the engine was allowed to stabilise for 7 minutes before recording and measurements represent average behaviour over 60 seconds. 4 conductances and 7 convective correlations).5%). These are shown with respect to the engine speed/torque map in figure 3 (a).1 Experimental Setup and Test Points A variable geometry turbocharger has been instrumented to capture fluid and structure T10mm temperatures at each of the locations described by the nodes of the model. The engine was e operated on an AC transient dynamometer. For the compressor and Tneu turbine housings. For all gas a) Gas temperature measurements in ducts temperatures. the aim of this work is to assess the predictive capability of the model for a new turbocharger. multiple thermocouples were used to capture distributions across the cross section of the ducts (figure 2a). Mass flow was measured at engine intake using an ABB Sensyflow hot wire flow meter (accuracy b) Compressor/Turbine <1%). Pressures were measured using Kistler Scroll Piezo-Resistive sensors (accuracy 0. based on the parameters established previously for a similar device. 3 APPROACH Tcentre T15mm 3. The turbocharger was installed on a i 2. gas and metal temperatures were measured using Figure 2: Thermal 1. 20 BMEP (bar) 20 Engine Transient 10 BMEP (bar) 15 Steady 4000 0 Speed (r/min) 10 5 2000 0 0 1000 2000 3000 4000 0 2000 4000 6000 Brake Speed (rpm) Time (s) (a) (b) Figure 3: Engine speed and torque operating points for stably and transient experiments (BMEP: Brake Mean Effective Pressure) The engine was operated over under thermally stable and transient conditions. Although an experimental procedure has been determined to characterise these parameters for an individual turbocharger. For steady state tests. temperatures were measured at 3 azimuths and 2 depths to provide E information about temperature distributions n Tamb g (figure 2b).2L Diesel engine for which the usual application Teng n is a light commercial vehicle. this replicating an industrial scenario.2 Model parameterisation The turbocharger thermal model contains a number of parameters that are not easily determined by inspection of the geometry without considerable simplification (5 capacitances.2. A range of turbocharger sizes has previously been measured and simple correlations have been observed between external geometries and thermal parameters: these have been used to calculate the parameters in this study. transient experiment consisted of a series of 32 step changes as shown in figure 3 (b) with a hold time of 3 minutes between steps. (6. the opening was tuned to give measured compressor mass flow at measured compressor speed. pressure ratio and mass flow predictions there is no significant change in the accuracy of the simulation by 107 . Figure 4: Turbocharger model boundary conditions for steady and transient simulations To avoid inaccuracies of a full engine model. this tests is superimposed onto figure 3 (a) in grey. The turbine map was determined from compressor enthalpy rise and is therefore adiabatic with respect to turbine performance. 16). 14). 3. however this includes mechanical losses and was modified by applying the mechanical losses models developed by Serrano et al. 4 RESULTS 4. 13) has been used to create the turbocharger model with the connecting ducts as shown in figure 4.1 Steady State Conditions The model accuracy with and without heat transfer simulation are compared in table 1. the PID approach is no longer possible as the dynamics of the controller would interfere with the physical dynamics. This approach was preferred to imposing compressor outlet pressure as it allows simulation errors to occur in pressure and mass flow. 2) The VGT position and compressor outlet orifice were fixed based on the results from 1. In terms of the turbocharger operating speed. On turbine side outlet pressure was imposed and VGT position was tuned to match measured mass flow at measured turbine speed. For the transient experiments.2 Model Control One-dimensional simulation code OpenWAMTM (12. A back-pressure valve has been introduced at compressor outlet to represent the flow restriction of the engine. measured conditions have been imposed at the compressor and turbine inlets (pulsating pressure for turbine). the turbocharger speed was now calculated from the compressor and turbine power balance. To avoid this problem look up tables were derived from the steady state data and interrogated via interpolation during transient simulations. The turbine is represented by a series of two nozzles with an intermediate reservoir to account for acoustic effects (15. This highlights the compressor and turbine models (that are based on the manufacturer’s maps) connected via a mechanical shaft and the lumped thermal model of the housing. For steady state the simulations were conducted in two phases: 1) The turbocharger was simulated at fixed speed whilst the flow coefficient of the orifice at compressor outlet and the turbine VGT positions were determined to match compressor and turbine flows respectively. 0% 5. For the housings. previous studies by Romagnoli and Martinez-Botas (5) have shown that the temperature can vary locally by more than the model accuracy when operated on- engine.5% 5. Turbine outlet temperature prediction is improved somewhat with the RMSE reducing from 29oC to 13oC whilst compressor outlet temperature prediction remains similar.0% Compressor Outlet Temperature 7oC 8.7% 2. with more than half this heat transferring to ambient. A number of possible sources for these errors are: engine installation effects.0% 2.3% 4.4oC 11oC 10oC Compressor Housing Temperature N/A 17oC N/A 12oC Turbine Pressure Ratio Imposed Turbine Mass Flow 2.1% 4.7% Compressor Mass Flow 2.3% Compressor pressure Ratio 3. The central bearing housing node (figure 5b) shows that higher heat flow from the turbine housing is primarily absorbed into the oil. At higher engine powers. Considering first node T (figure 5a).2% 4. The thermal model allows a number of additional temperatures to be estimated: notably the oil temperature rise in the turbocharger and housing temperatures to within 10-17oC.3% 2.1% 4. therefore discrepancies may be a result thermocouple installation. operation under pulsating flows and manufacturing tolerances. Figure 6 shows the proportions of temperature 108 . the heat flow to the compressor housing reduces which is also highlighted by the reversal of heat flow between air and compressor housing shown in figure 5c.1% 6. Table 1: Root Mean Square Error (RMSE) for steady and transient simulations. small errors in the operating maps will be compounded as they will shift simulated speed. heat flows are presented for each node against engine operating point.adding the thermal model. the total heat transfer from exhaust gases increases with torque and power.3% Turbine Outlet Temperature 29oC 13oC 35oC 14oC o Turbine Housing Temperature N/A 13 C N/A 15oC Oil Outlet temperature N/A 9oC N/A 7oC Bearing Housing Temperature N/A 11oC N/A 16oC 2500 900 250 Q (T to AMB) Q (BH to Oil) Q (C to Amb) 800 200 2000 Q (T to H1) Q (BH to AMB) Q (C to A) 700 Q (BH to C) 150 600 Heat (W) Heat (W) Heat (W) 1500 100 500 400 50 1000 300 0 500 200 100 -50 0 0 -100 1000rpm/50Nm 1500rpm/50Nm 2000rpm/50Nm 2500rpm/50Nm 3000rpm/50Nm 1000rpm/100Nm 2000rpm/100Nm 1000rpm/150Nm 3000rpm/150Nm 1500rpm/200Nm 2500rpm/200Nm 1500rpm/300Nm 2000rpm/300Nm 2500rpm/300Nm 1000rpm/100Nm 2000rpm/100Nm 1000rpm/150Nm 3000rpm/150Nm 1500rpm/200Nm 2500rpm/200Nm 1500rpm/300Nm 2000rpm/300Nm 2500rpm/300Nm 1000rpm/50Nm 1500rpm/50Nm 2000rpm/50Nm 2500rpm/50Nm 3000rpm/50Nm 1000rpm/50Nm 1500rpm/50Nm 2000rpm/50Nm 2500rpm/50Nm 3000rpm/50Nm 1000rpm/100Nm 2000rpm/100Nm 1000rpm/150Nm 3000rpm/150Nm 1500rpm/200Nm 2500rpm/200Nm 1500rpm/300Nm 2000rpm/300Nm 2500rpm/300Nm (a) Node T (b) Nodes H1/H2/H3 (c) Node C Figure 5: Predicted heat flows in selected turbocharger nodes A breakdown of thermal energy flows between selected model nodes is given in figure 5. ranked as increasing engine power. with and without heat transfer (HT) model Steady State Transient State Factor HT OFF HT ON HT OFF HT ON Turbocharger Speed 4. mass flow and pressure ratio. The oil temperature rise Includes both friction and forced convective heat transfers. During the simulation.1% 6.4% 3. especially at high engine powers (with high compression/expansion ratios). but that the opposite is not true. Internal Conduction controls the amount of heat conducting between the turbine and compressor housing. the following observations are made: Turbine Internal convection has a direct effect on the heat flux between the gas and the housing (20-30%). The study shows that the turbine behaviour is independent of compressor heat transfer. Heat flow through nodes H1. No significant effect on turbine temperatures was observed. External Heat Transfer had a significant influence on all four temperatures with gas temperatures and metal temperature varying ±2-14oC and ±5-80oC respectively. This is not directly proportional to heat transfer coefficient because node T varies ±20-80oC whilst outlet gas temperature varied up to ±8oC.change attributable to work and heat transfer: it is therefore unrealistic to assume that all simulation errors could be corrected by the heat transfer model. Table 2: Ranges of sensitivity tested Factor Variable Variation Turbine Internal convection hAGAS/T (eq 2) -50% +50% Compressor internal convection hAC/AIR (eq 2) -50% +50% External Heat transfer hAX/AMB (eq 2) 0 +100% Internal Conduction KX/Y (eq 1) -50% +50% Compressor Internal Convection caused variation in compressor outlet gas temperature and compressor housing of up to ±4oC. H2 and H3 is affected causing changes in node C of ±2oC. The crucial influence of internal conduction for estimating metal temperatures demonstrates the need for a rigorous identification whilst the significant effect of external heat transfer highlights the need for accurate control and modelling of this. 109 . Based on the results in figure 7. The sensitivity showed a ±19oC and ±7oC variation in turbine and compressor housing respectively and ±4oC and ±3oC in compressor and turbine gas outlet temperature respectively. Metal temperatures are more sensitive to model parameters than gas temperatures. 675 200 625 Heat 180 Heat 575 Expansion 160 Compression 525 140 Tgas (oC) Tgas (oC) 475 120 425 100 375 80 325 60 275 40 225 20 100rpm/1000Nm 100rpm/2000Nm 150rpm/1000Nm 150rpm/3000Nm 200rpm/1500Nm 200rpm/2500Nm 300rpm/1500Nm 300rpm/2000Nm 300rpm/2500Nm 50rpm/1000Nm 50rpm/1500Nm 50rpm/2000Nm 50rpm/2500Nm 50rpm/3000Nm 50rpm/1000Nm 50rpm/1500Nm 50rpm/2000Nm 50rpm/2500Nm 50rpm/3000Nm 100rpm/1000Nm 100rpm/2000Nm 150rpm/1000Nm 150rpm/3000Nm 200rpm/1500Nm 200rpm/2500Nm 300rpm/1500Nm 300rpm/2000Nm 300rpm/2500Nm (a) Turbine (b) Compressor Figure 6: Temperature changes due to compression/expansion and heat transfer for (a) turbine and (b) compressor A sensitivity study was conducted to quantify the influence of each parameter in Table 2. the use here of the heat transfer model using parameters with high uncertainty has shown improved prediction of turbine outlet gas temperature. Nevertheless. but less than 1oC change in compressor outlet temperature. (c) Node T and (d) Node C (Axis refers to different operating conditions. on the compressor theses were ±7oC and ±4oC respectively. 550 170 500 Tgas comp out (oC) Tgas turb out (oC) 150 450 130 400 350 110 300 90 250 70 200 50 1 2 3 4 5 1 2 3 4 5 (a) (b) 620 180 520 160 T node C (oC) T node T (oC) 140 420 120 320 100 220 80 120 60 1 2 3 4 5 1 2 3 4 5 (c) (d) 1 – 50Nm/1000rpm 2 – 50Nm/2500rpm 3 – 100Nm/2000rpm 4 – 300Nm/2500rpm 5 – 300Nm/1500rpm Figure 7: Model sensitivity to model parameters for (a) turbine gas outlet. A sensitivity analysis has also been carried out where thermal capacity was varied ±50%: figure 8 shows the results over a particular step in engine operating point from 280Nm/2000rpm to 60Nm/2300rpm. Overall the prediction is of similar accuracy to steady state performance. ±50% in housing thermal capacity resulted in ±20-40oC and ±8-11oC in metal and gas temperature prediction over the transient. however under these conditions in addition to the parameters discussed above. On the turbine side. As the engine power is reduced. (b) compressor gas outlet. there is a rapid drop in compressor and turbine outlet temperature associated with the expansion and compression processes. HT: Heat Transfer) 4. associated with heat transfer and thermal inertia. 110 . the thermal capacity of each node is also important.2 Transient conditions The prediction performance of the “best estimate” model over transient events is given in table 1 and compared graphically to measured results in figure 8. Subsequently a much slower temperature reduction is observed. M. and A. vol. and A. REFERENCES [1] M. Experimental Study of the Turbine Inlet Gas Temperature Influence on Turbocharger Performance. D. 55. 38. vol. 58-77. P. pp. 2004. L. 2012. SAE Paper Number 2006-01-0023. Maiboom. Serrano. SAE Paper Number 2007-01-1559. This showed that turbine outlet temperature prediction RMSE could be reduced from 29 to 13oC for steady state conditions whilst other predictions were of similar accuracy. and A. 111 . Dris. Serrano. Tiseira. pp. A. Heat Transfer Analysis in a Turbocharger Compressor: Modeling and Experiments. F. Martinez-Botas. R. J. Heat transfer analysis in a turbocharger turbine: An experimental and computational evaluation. 132(4). C. A. [2] J. 2007. Garcia-Cuevas. V. [4] N. 888-898. Hetet. R. Journal of Engineering for Gas Turbines and Power. 2006. Olmeda. The analysis of heat transfer in automotive turbochargers. Shaaban. Tiseira. [6] J. K. Under transient events. Baines. 2013. Experimental investigation and extended simulation of turbocharger non-adiabatic performance. Wygant. Chesse. The model allowed metal temperature to be estimated to within 17oC and 15oC for compressor and turbine housings respectively. However the sensitivity study has suggested that a further accuracy improvement would be expected if a full thermal characterisation was undertaken. Cormerais. A sensitivity study showed that uncertainty in the model parameters could affect gas temperatures up to ±4oC and metal temperatures up to 80oC. vol. Romagnoli and R. 2010. Applied Thermal Engineering. the uncertainty associated with thermal capacity of the nodes increases the errors but this has limited influence on gas temperatures. (a) (b) Figure 8: Model sensitivity to thermal capacitance for (a) turbine and (b) compressor 5 CONCLUSIONS A study of the predictive performance of a turbocharger heat transfer model has been undertaken under steady and transient conditions. The results show that at this stage it is possible to improve the prediction of turbine outlet temperature on-engine by using the heat transfer model and inferring properties from a similar device. PhD. Universität Hannover. [5] A. Lefebvre. Energy. Dolz. Cervelló. P. Fachbereich Maschinenbau. Guardiola. Theoretical and experimental study of mechanical losses inautomotive turbochargers. [3] S. and C. García-Cuevas. Dolz. [12] OpenWAM. R. Galindo. Measurement Science & Technology.-F. R. F. vol. R. vol. PhD. pp. Available: www. 2013. Smith. vol. Linnhoff. J. R. Reyes-Belmonte.org [13] J. Turbocharger heat transfer modeling under steady and transient conditions. Cervelló. 112 . Serrano. pp. Serrano. Energy Conversion and Management. pp. [10] S. Journal of Engineering for Gas Turbines and Power- Transactions of the Asme. Engines.May 18. and F. 49. vol. Pucher. and J. J. 2006 . Description of a Semi- Independent Time Discretization Methodology for a One-Dimensional Gas Dynamics Model. Lefebvre. Reyes-Belmonte. Olmeda. J. Piqueras. 131. M. V. London. 57. International Journal of Thermodynamics. Vidal. P.and one- dimensional gas dynamics codes for internal combustion engines modelling. Determination of heat flows inside turbochargers by means of a one dimensional lumped model. 193-202. A. and A. pp. Arnau. Olmeda. May 2009. Seume. vol. J. Shaaban. 2013. Mar 2010. Tiseira. A. and M. Serrano. Submitted to Energy. [15] J. Arnau. 2013. Tiseira. 2006. Hetet. Chesse. A. 729-738. 2006. [16] M. Mathematical and Computer Modelling. F. L. 1847-1852. A. 21. 2009. A. A. J. An experimental procedure to determine heat transfer properties of turbochargers. F. vol. Cormerais. May 17. F. United Kingdom. P. 12. [11] F. presented at the 8th International Conference on Turbochargers and Turbocharging. V. A model of turbocharger radial turbines appropriate to be used in zero. 2008. Dombrovsky. Dolz. R. R. and P. Berndt. and C. Part Load performance prediction of turbocharged engines. P. Universitat Politècnica de València. Departamento de Máquinas y Motores Térmicos. Olmeda. Payri. Olmeda. 2013. Paez. SAE Int. [9] J. External heat losses in small turbocharger: model and experiments. P. and L. [14] J. 3729-3745. València. and H.Openwam. Contribution to the Experimental Characterization and 1-D Modelling of Turbochargers for IC Engines. J. 2012. Arnau. J. H. J. 6. [8] P. Serrano.[7] M. Arnau. Importance of Mechanical Losses Modeling in the Performance Prediction of Radial Turbochargers under Pulsating Flow Conditions. UK ABSTRACT This paper describes a novel method of designing the HP compressor of a two-stage turbocharger using a 1-D code coupled with an optimisation algorithm.A 1-D analytical code for the design and multi-objective optimisation of high- pressure compressors within two-stage turbochargers for marine applications O F Okhuahesogie1. Subscripts pressure (J/kgK) 1 Inducer inlet location Shaft speed (rev/min) 2 Impeller exit location Mach Number 3 Diffuser exit location Number of main blades Blade Total number of blades Flow Blade velocity (m/s) ℎ Thickness Gas absolute velocity (m/s) Slip Gas relative velocity (m/s) − Total to static Gas absolute angle to axial − Total to total direction ℎ Hub Gas relative angle (deg) Shroud Blade angle (deg) Choke Slip factor Average Work input coefficient Relative _______________________________________ © The author(s) and/or their employer(s). NOTATION Pressure (bar) Local speed of sound (m/s) Temperature (K) Blade height (m) Area ( ) Velocity correction coeff. The 1-D code solves relevant thermodynamic and fluid dynamics equations given geometric variables. A multi- objective optimiser is used to return potential designs that will maximise compressor efficiency and minimise deviation from the target mass flow. Efficiency Offset constant Density ( ) ζ Entropy loss coeff. J Stewart1. Number of cells in grid Ratio of specific heats Gas incidence angle(deg) Specific gas constant/Radius Inducer throat area (as used) Pi Mass flow rate (kg/s) Specific heat at const. M J W Riley1. P Roach2 1 School of Engineering. An example based on a HP compressor impeller and vaneless diffuser is presented to demonstrate the potential of this 1-D methodology. UK 2 Napier Turbochargers Ltd Lincoln. F Heyes2. These designs are then taken forward for detailed 3D-CFD analysis. 2014 125 . University of Lincoln. the target pressure ratio and the operating speed as inputs. how the optimisation task has been phrased for a Differential Evolution for Multi-Objective Optimisation (DEMO). It is difficult to achieve high pressure ratios in single stage compression due to limits on material strength (e. The current work describes the configuration of the 1-D code. a 1-D code is used to calculate centrifugal compressor performance based on a known parametric geometry at a desired operating point using empirical correlations without looking into the details of the flow within the components. S. the preliminary design phase of centrifugal compressors often applies a one-dimensional approach to solve the thermodynamics and fluid equations that describe the internal flow physics. Differential Evolution (5) has attained much popularity in the optimisation community. In the direct method.g. This method computes performance parameters such as stage pressure ratio and efficiency for a given geometry and inlet conditions. aluminium) and the prohibitive cost of more durable materials (e.g.1 Design optimisation process Most turbomachinery design processes start from simple 1-D calculations of a desired operating point and compressor characteristics before a more detailed design 3D analysis is carried out. DEMO. 2. However. and presents the results using a test case of the preliminary design phase of a high pressure centrifugal compressor within a two-stage turbocharger. The inverse design algorithm implemented by Yuri Baba et al (3) calculates the required one- dimensional geometry and relies on a trusted database of loss models. This computationally inexpensive method gives results that guide the engineer towards the optimum design space for further exploration with 3-D CFD. when used in gas turbines). performance is calculated from a given compressor geometry. was found to have comparable performance to other state-of-the-art optimisation algorithms (6). when a single stage compressor is designed for pressure ratios of up to 9 (e. particularly when the design process relies heavily on the use of computationally. The compression ratio achievable over the past 40 years using aluminium for single stage compression has reached a limit of about 6:1 (2) which is not high enough to meet IMO III requirements as pressure ratios above 8:1 are more suitable. The design process discussed in this paper focuses on 1-D aerodynamic optimisation of the design point of a centrifugal compressor. More so. The development of new.1 INTRODUCTION International Maritime Organisation (IMO) legislation demands a reduction in NOx emissions to 20% of the pre 2011 levels by 2016. the two potentially conflicting objectives are to maximise the efficiency 126 . Yuri Biba and Peter Menegay (3) identified two main methods used in the preliminary design process for centrifugal compressors. For the inverse design approach. High pressure-ratio turbocharging can be achieved using two-stage turbochargers.and time-expensive CFD models. 2 CURRENT STUDY In this study. For this work. Kamaleshaiah et al (4) implemented a centrifugal compressor performance algorithm based on a direct prediction method with empiricism for the loss models and boundary layer growth within the blade channel. The variant used here. the geometry is calculated to meet a required performance. The development of new turbochargers is challenging in a highly competitive market place. titanium) (1).g. high pressure-ratio turbochargers is the key enabling technology that will allow marine engines to meet this target. M. the efficiency and map width are significantly lowered and not sufficient for turbocharger applications. A multi-objective optimisation algorithm calls this 1-D code and seeks the most promising candidates by modifying the parameters that specify the geometry. makes it feasible for use with multi-objective optimisation algorithms on a regular PC configuration. Choke mass flow corrected with a blockage factor (7) is calculated using equation 1.1 Inducer calculations At the inducer. The program then calculates the mass flow and efficiency based on an initial geometry.and minimise the deviation from the desired mass flow at the target design speed line and pressure ratio as shown in figure 1. The design objective of the 1-D algorithm is to find the geometry of a compressor that will deliver a required pressure ratio at the design speed line. Gas and blade velocity components are defined at the root-mean-square location.3. It solves the thermodynamic and fluid dynamics equations in order to describe the flow physics in a radial compressor impeller and vaneless channel diffuser.2 Description of computer program and algorithm The 1-D code used in this paper was written in C++ for the Linux operating system. The algorithm uses a number of iteration loops until error is within 1% as shown in figure 2. the geometric and thermodynamic properties are calculated using the method described by Whitfield A et al (7). 2 −1 = 1+ (1) +1 2 127 .5 GHz. The speed of execution of the code (1 second) on an Intel i5-2520M 2.3 Compressor mathematical modelling 2. The user can later modify the geometry manually or automatically using optimisation algorithm described here until a suitable mass flow and efficiency is achieved. Figure 1: Design Optimisation Flow Chart Figure 2: Algorithm Flow Chart 2. 2. Inducer blockage is then updated using blockage correlations as described in section 2.3. =1− (3) 2 ( − ) The peak efficiency is corrected to obtain the total-to-total efficiency using correlations described in section 2. The velocity components and thermodynamic properties at the discharge are then calculated using the equations as described by Whitfield et al (7). J (1967) in (7.3. Slip factor is related to slip velocity and blade speed as shown in equation 2. The 1-D code presented here allows the user/DEMO to select any of the three slip correlations or a no-slip option. =1− (2) Efficiency is calculated using an equation for peak total-to-total efficiency via equation 3 where is the entropy or energy loss coefficient (10). for example those described by Stodola A.4. 2.3.D (1952a) and Wiesner F. an initial estimate of mass flow (which must be less than choke mass flow) and static density is used to start the iterative calculation in the inducer.2 Discharge calculations A typical velocity triangle at the exit of an impeller with backward swept blade angle (backsweep angle) is shown in figure 3. can be calculated using established correlations. The other thermodynamic and velocity parameters are then computed. Slip factor . Stanitz J.3 Diffuser calculations A mathematical model for one-dimensional flow analysis for vaneless diffuser which accounts for heat transfer effects and wall friction effect in both the linear and angular momentum equations (11) is implemented in the 1-D code using equations 4 to 8.At this point. the process continues until the error is less than 1%. which estimates the deviation of the gas angle from the blade angle. 1 : + − + =0 (4) 128 . 2.4 and the density loop is iterated until the error is less than 1%. hence the calculation is started by using an initial estimate of impeller exit density ( ). Figure 3: Impeller discharge velocity triangle The dashed lines (blue and green) represent gas absolute velocity and gas relative velocity. The estimated density is corrected using the calculated value. These dash lines are as a result of slip (relative flow at impeller exit not tangential to the blade profile) (8).9). the gas absolute ( ) and relative ( ) flow angles are not known. At the discharge.3. (1927). The black and red lines represent blade and slip velocity respectively. Friction factor can be expressed in terms of Reynolds number (7).4. This variation affects the location where separation may occur in the impeller. validation calculations carried out using angle offset correlations shows reasonable correlation with CFD results.3. 2.2 Axial Length – Discharge Diameter Ratio: The impeller length affects the length of the compressor. The correlation is expressed in equation 11.4.3. 2. However. impeller bore stresses and aerodynamic performance (13). 10) Equations 4 to 10 were modelled using an upwind finite difference scheme of first order accuracy (12) on a grid of 40 steps. the more the flow will turn from inducer to discharge.3. The calculated static pressure at diffuser exit is compared with the target static pressure and mass flow is varied until the error is less than 0.4. Birdi (14) suggested a correlation of the ratio of axial length to discharge diameter for a range of inlet Mach number according to equation 12. 1 1 1 1 − − + − + = 0 = (9.1 Angle offset: This is based on the idea that the larger the difference between inducer and discharge blade angles.1 Friction loss Friction loss in the diffuser is implemented via a friction factor in the radial and tangential momentum equations describing diffuser flow physics.2.4.2.25%.3. 129 . axial length – discharge diameter ratio and velocity coefficient) as described below.4 Loss models correlations 2. The calculated value is corrected until the error is less than 1% to end the efficiency loop.3.2. the shaft dynamics. The velocity and thermodynamic parameters at the exit of the diffuser are determined at the end of the calculation. : + + =0 (5) 1 1 1 1 : + + + =0 (6) −1 (7. Larger turning increases flow separation and losses due to blockage.3. 2. It is therefore necessary to have a balance of axial length and discharge diameter. The target total-to-total efficiency can then be calculated using target values and compared with the calculated total-to-total efficiency described in section 2. 8) ∶ = : = + 2 Equations 4 to 8 can be re-arranged to obtain the forms below. =1− × − (11) ℎ 0 ≤ ≤1 The limitation of this idea is that the impeller is a 3-D object and blade angles vary from inducer to discharge. 2.2 Impeller blockage Blockage correlations developed and implemented in the code are based on the concepts (angle offset. 2. The pressure loss during each step is formulated with the equation of a straight line using equation 17. The concept is borrowed from the opening angle correlation of a volute exit cone where pressure loss factor varies from a value of 0.37 for inlet Mach numbers of 0.3 Velocity coefficient: According to Blasius theory for flat plate incompressible flow (15).382 ≈ ≈ (14. In this 1-D code.2.2 for typical ranges of diameter ratios.017 × 2 − 0. The effect of on impeller blockage was 2 2 estimated using equation 13 ≤ ≤ ∶ = 1 .85. 4. < 1 . ∆ = (0. the equation above yields values of 0.4. > 1.3 Opening Loss An opening loss model is implemented in the diffuser to predict pressure drops due to rapid increase in diffuser area from inlet to exit. boundary layer grows from inlet to outlet and is inversely proportional to Reynolds number or velocity over distance downstream of the boundary layer as shown in equations 14 and 15.28 and = 0. ≤ ≤ .91 0.37.3. 130 . hence a model to account for the effect of gas relative velocity on impeller blockage is modelled with equation 16 where is a velocity correlation coefficient.15 for an opening angle of 100 to 1 for an opening angle of 600 (8).32 = 0.35 ≤ ≤ 0.02)/ (17) The opening angle 2 is calculated by first estimating equivalent radius for diffuser inlet and outlet assuming the diffuser was a cone using equations 18 and 19. otherwise is set to 1 in the code and choking will cause blockage when . According to Birdi. = . + 1− where = 0.9 – 1. < _ ∶ = (13) _ . In this 1-D code. (16) The above equation is only valid when . where = 0.3. The overall impeller blockage factor is computed as a product of all the blockage factors. 15) where is the distance downstream of the start of the boundary layer and is the boundary layer thickness. . _ > _ ∶ = . Reasonable values are in the range 0.4. − (12) = .8. 2.32 – 0. . Reynolds number is directly proportional to the gas relative velocity through the impeller blades. 95 (21) = 1 ∶ < 0.961 2 0.000 cells) and experimental data of the same geometry as shown in table 1. = ∶ > 0 (22) =1∶ < 0 The actual total-to-total efficiency is then calculated by multiplying the peak efficiency with the corrections.95 . 19) − The ratio of diffuser outlet to inlet height is included so it has a significant influence on the opening angle (7). 2.3.827 Experiment 9.609 1-D Code (Stanitz) 8. the rate of diffusion in the impeller using limits suggested in (8) and the gas incidence angle at the inducer.900 2 0. however. As a result.725 ANSYS CFD 8. Surge point in the 1-D code was determined to be when the iteration does not converge or when the gas angle at impeller or diffuser exit exceeds 80o depending on which situation occurs first (8). − = = = tan (18. = 1 ∶ 65 ≤ ∝ ≤ 80 (20) . the design point geometry is analysed using the 1-D code for different slip correlations and compared with ANSYS CFX CFD results (on a mesh of about 600. Loss models were then developed according to the equations 20 to 22.4 Validation of the 1-D code The 1-D code was originally written to calculate a design point on a compressor map.105 2 0.790 131 .4. it can be used to produce an approximate compressor map by running the code at a different outlet pressure at the same speed. Table 1: Design points comparison Mass Flow Pressure Total-static Corrected Ratio efficiency 1-D Code (No Slip) 7. 2.993 2 0.3.728 1-D Code (Wiesner) 9.2 is corrected using the impeller exit gas angle. = ∶ ≥ 0.4 Efficiency Correlations Peak efficiency calculated in section 2.702 1-D Code (Stodola) 8. = ∝ : ∝ < 65 ∝ > 80 . due to its functionality.500 2 0.212 2 0. 1 Optimised design(s) The current design of the HP compressor is set to operate at a pressure ratio of 2:1 and a target mass flow of about 11. The performance curve in figure 4 shows that the 1-D code using Stanitz and Stodola (7) slip correlations predicts efficiency to within 4% of CFD/experimental values towards the surge region and to a maximum of around 20% towards choke.60 CFD_Old Exp 1. New designs were selected based on efficiency and closeness to target mass flow. the 1-D code was able to predict choke more accurately compared to surge. Impeller diffusion ratio ( ) must be between 0. Any differences between the 1-D and CFD results are due to extensive use of correlations to capture certain 3D physics in the 1-D model. the target mass flow was 11.80 D0_Wiesner Exp D0_Stanitz 1. Also.95 Diffuser Opening Angle must be less than 15 degrees Minimise deviation of calculated mass flow from target mass flow(objective) Maximise compressor efficiency (objective) 3.25 Impeller exit gas angle must be between 65 and 80 degrees.4kgs-1. The optimiser was set to reject candidate geometries with mass flows not within 20% of target while finding a design as close as possible to the target mass flow.25 and 0. In table 2. D0 refers to original design.20 D0_NoSlip D0_Stodola D0_Wiesner 2. 3 RESULTS DEMO (16) was set up with 75 candidates in a generation and run nominally for 1000 generations.80 Pressure Ratio 2. 2.4kgs-1 at 66.00 D0_Stanitz D0_NoSlip D0_Stodola 0. Overall.31 rads-1 and has an inlet stagnation pressure and temperature of 350000 Pa and 333K respectively. The 1-D model was able to predict the width of the performance curve to within 20% deviation from CFD/experiment measurements. the result shows that the correlations developed here are helpful in capturing the basic physics. the following criteria were set and penalised if violated: Impeller exit blade speed must be less than 580 m/s Relative Mach number at inducer shroud must be less than 1. b): Validation Performance Curves Table 1 shows that the 1-D code (except no slip correlation) under-estimation of mass flow is within 10% of CFD and experimental values at design point.60 CFD_old 1. 132 . However. The most obvious deviation in cases presented is the no-slip correlation model and as such will not be an option in the optimisation calculations. by the 600th generation it was clear that the progression of the Pareto front shown in figure 5 had stagnated as the objective functions ceased to improve.40 Total-to-static efficienccy 0. There is more deviation in between the 1-D code and CFD/experiment in estimated efficiency.40 0. During the optimisation.40 4 6 8 10 4 5 6 7 8 9 10 11 Corrected Mass Flow Corrected Mass Flow Figure 4 (a. For this case. Table 2 shows the current design and 5 (2 from the 600th generation and 3 from the 300th generation) new possible improved designs using the 1-D code. Based on figures 6a and 6b.9 D0_Stanitz CFD CFD Exp Exp D1 D1 1. 0. All of the new designs have fairly similar performance.8 Gen_100 Gen_200 Gen_300 0.8 Gen_400 Gen_500 Gen_600 0.20 3 5 7 9 11 3 5 7 9 11 Corrected Mass Flow Corrected Mass Flow Figure 6 (a.80 Pressure Ratio 2.7 0. there is a reasonable distance between the peak efficiency mass flow and choke mass flow in design D2 compared to the other designs.7 0 5 10 15 Mass Flow Violation % Figure 5: Pareto front of generations Table 2: 1-D results of original and new designs Figures 6a and 6b compare the 1-D performance curve and speed line of the new and existing design with CFD and experimental data.9 Total-Static Efficiency Gen_10 0. 1.40 D2 D2 D3 D3 D4 D4 D5 D5 1. but design D2 gives a reasonable operating range and higher pressure (about 5% more) near surge compared to the other new designs.5 0.3 Total-to-static Efficiency 0. design D2 was chosen as the best compromise design. b): Performance curve of new and old designs 133 .00 2. More so.60 D0_Stanitz 1.1 D0_Stodola D0_Stodola D0_Wiesner D0_Wiesner 0. 4 D2_1D 0. Also.80 0. Validation of the 1D code with a 3D CFD solver and experimental data revealed that reasonable approximations for the pressure ratio and efficiency were being returned with a greatly reduced computational cost. pressure ratio and shaft speed with constraints. Emissions – A new challenge for turbocharging. Paper No. KTP Associate Conference.50 0. Stewart J. Vienna. G. 6 REFERENCE LIST [1] Okhuahesogie O. this work confirms that such a method can be used to calculate an initial feasible design in the early stages of the design process of a new compressor.6 2. Mathery C. KTP is funded by the Technology Strategy Board along with the other government funded organisations. Although the equations used to model the 1-D system do not capture the full physics of the compressor flow.60 1. technology and skills that reside within the UK Knowledge Base. 245. This would ultimately speed up the design process if less computationally expensive methods described here are used in the early phase of the design process by reducing the design space to explore during the more computationally expensive CFD phase..00 2.00 Pressure Ratio Radius 0. 2012 [2] Codan E. CIMAC. The code has been linked with a multi-objective optimiser and ran hundreds of potential geometries to find the compressor geometry with optimum efficiency at a target mass flow. the new design gives slightly higher efficiencies at lower pressures. J.5 1 4 6 8 10 4 6 8 10 Axial Length Corrected Mass Flow Corrected Mass Flow Figure 7 (a. Brighton. 1 1. c): Meridional view and performance curves of new and old designs Figure 7 shows good correlation in mass flow calculations at different outlet pressures. Design Optimization of a two-stage Compressor Impeller. b. 4 CONCLUSIONS A computer code has been developed to predict the performance of a centrifugal compressor at the design point.E. Any difference in efficiency estimation is due to the effect of blade curvature on efficiency which is not captured in the 1-D model.2 D2_CFD D2_CFD CFD_Original CFD_Original 0 1. The code is shown to be accurate enough to be relied upon during the preliminary design phase of a centrifugal compressor. Heyes F.25 0.F.75 D2_1D 0. The code can also be used to predict the performance at off-design point by changing the outlet pressure at a specific speed for a given geometry. Roach P. 2007 134 .40 0 0. KTP aims to help businesses to improve their competitiveness and productivity through the better use of knowledge.8 D0 Pressure Ratio D2 0. 5 ACKNOWLEDGEMENT This partnership received financial support from the Knowledge Transfer Partnerships programme (KTP). A simple and efficient adaptive scheme for global optimization over continuous spaces. 213. Robic. Slovenia 135 . Engineering Education System.. Tusar.A. October 1967 [10] Denton J. 10:75-84. Vol. Robinson C J. Centrifugal Compressor Design. including effects of friction. Mixed flow compressors. 2000 [13] Came P. 1.S.[3] Yuri B. SI-1000 Ljubljana. 1999 [14] Birdi K. Indian Institute of Technology. John Wiley and Sons. S. N. von Karman Institute for Fluid Dynamics. Jozef Stefan Institutez. Centrifugal Compressors Analysis and Design. Journal of Engineering for Power. 4th Edition. Part C. Hoffmann. No4. Mech. Cranfield University Short Course on Centrifugal Compressors. Department of Intelligent Systems.. Vol. Madras [5] R. Department of Intelligent Systems. Jozef Stefan Institute. Slovenia [7] Whitfiled A. Chiang Computational Fluid Dynamics. Design of Radial Turbomachines. Storn. International Journal of Rotating Machinery. van den Braembussche.. Price Differential Evolution . P Menegay Inverse Design of Centrifugal Compressor Stages Using a Meanline Approach. Boundary layers in fluids with little friction.T. 89. Ramamurthy An Improved Method of Centrifugal Compressor Performance Prediction. Engrs. pp 558-572. NACA TN 2610. Loss Mechanisms in Turbomachines. heat transfer and area change. Vol. Baines N. Jamova 39. B. One-dimensional compressible flow in vaneless diffusers of radial or mixed-flow centrifugal compressors. M. Filipic DEMO: Differential Evolution for Multiobjective Optimization.D. A. Instn. Berkeley. F. Venkatrayulu. Kamaleshaiah. 93-GT-435. Filipic Differential Evolution Versus Genetic Algorithms in Multi- objective Optimization. Proc. 1992 [15] H. K. ASME. International Computer Science Institute. 2004 [4] M. 1990 [8] R.C. February 2012 [9] Wiesner. Jamova 39. 1952 [12] K. 1993 [11] Stanitz J.J. CA 94704 [6] T. 1950 [16] T. Blasius. A review of slip factor for centrifugal impellers. 1947 Center Street. NACA technical memorandum 1256. May. B. SI- 1000 Ljubljana. S.. D. Course Note 192. P Davies. flexible multi-body dynamics and thermo-mechanical simulations and how they can be linked together to analyse a complete system. allowing higher EGR rate.1 Impact on Gasoline engine turbocharger control strategy Explanations are given below as to why modern gasoline turbocharging largely rely on fixed geometry / by-pass control of boost: _______________________________________ © The author(s) and/or their employer(s). This paper describes key elements of the kinematic tools developed by Honeywell Turbocharger Technologies in order to optimize turbocharger control solutions. Czech Republic ABSTRACT Over the past years. N Morand. noise and durability L Toussaint. the new emission regulation EU6 drives toward very low emission levels. M Marques. This is a strong trend. fine. This requires us to apply a System Engineering approach and improve our understanding of the behaviour of the entire control chain so that we can ultimately influence the design and drive performance at the System level. controllability. D Vlachy. This examples show how kinematic tools can be applied to achieve controllability and durability targets while accelerating development cycle time. R Mrazek Honeywell. Significant upgrade of chassis and powertrain are necessary to meet these challenging targets. France F Tomanec. 1. At the same time. It includes details on CAE applications such as fluid dynamics. the evolution of emission standards as well as the advent of Fuel economy targets has driven the need for higher controllability of boost pressure. 2014 137 . Turbocharging boosting system plays a major role in powertrain optimization. 1 INTRODUCTION Effective in 2014. for both emission control. accurate and sustainable control of boost is challenging the way mechatronics components are engineered. C Groves Honeywell. which will continue to be true for both gasoline and diesel engine applications as we prepare for further evolutions of standards and introduction of fuel economy regulations. While engine boundary conditions have become harsher. M Zatko. and fuel economy. fleet average C02 emission have to meet 130g/km. It highlights the importance of understanding the behaviour of the whole control chain from the actuator command to the turbocharger and engine response.Improvement of a turbocharger by-pass valve and impact on performance. The paper presents results from one main examples on a new waste-gate concept for Fixed geometry turbochargers primarily for gasoline applications. valve response time. thus increasing engine efficiency at key fuel economy in the point driving cycle. This is particularly key as a quick and high boost response is needed in case of high driver torque demand in order to have the downsized engine still behave like a bigger naturally aspirated engine. vibrations) in which the valve has to operate. the relatively low amount of air required by gasoline engines in typical rated power conditions still allows the engine to reach very high specific power outputs in the range of 90 to 120 kW/l.2 Impact on Diesel engine turbocharger control strategy By-pass controlled turbochargers also exist on passenger vehicle diesel applications but are becoming less and less used in developed countries as they are becoming non competitive in terms of specific power (<50 kW/l). As a consequence. orders of magnitudes are given for the key environmental conditions (temperature. First of all. thus allows the turbocharger to be controlled and not to produce too much boost pressure. gasoline engines typically do not require any boost pressure to operate in part load. pressure.The sizing of the by-pass also allows the boost to be reduced in certain operating conditions. There is no need to operate at low or zero boost as is the case with gasoline engines. an example of development of a new by-pass system is given as an illustration of kinematic simulation tools that have been developed in order to address arising concerns. despite the small turbocharger matching generally used to get good vehicle driveability.m/l & 75kW/l for a CO2 emission in the range of 85 to 120g/km in ECE fuel economy cycle. 138 .Heavy EGR rates required to pass emission standards even increases the compressor power needed in part load operating points.... 2 NEW WASTE-GATE CONCEPTS FOR FIXED GEOMETRY TURBOCHARGERS PRIMARILY FOR GASOLINE APPLICATIONS In the following section. even in the case of a small turbocharger. .Note finally that. This directly impacts the capability to control the driver’s torque demand without having to use the intake throttle and avoids generating an undesirable engine back pressure. improving as well low end transient performance. .Adopting a by-pass strategy also allows sizing the turbocharger in order to maximise boost at low engine speed. . Variable geometry turbochargers are already main stream for Diesel passenger vehicles and are becoming increasingly popular in Euro 6 applications for the following reasons: .With the stoichiometric burn. noise or durability with component level targets (closed valve sealing.Allows reaching higher and higher specific torque and power while providing world class fuel economy: Vehicles reaching up to 200N. As such. In a later portion. . most of the typical control strategies for passenger vehicle diesels rely on variable geometry. linking OEM requirements like performance. 1. a by-pass system facilitates the regulation of exhaust gas energy available at the turbine wheel and.). a qualitative analysis of key by-pass functionalities is provided. in particular on the low load steady state points of fuel economy driving cycles.The high air to fuel ratio in Diesel engines requires a significantly higher boost demand in most engine operating conditions. . The key customer impacted criteria is therefore here vehicle driveability (comfort)..Enable quick valve response on transient event requiring large valve travel from either large open to close or close to open. This condition is typically required to either achieve low end torque at low engine speed (< 1300 RPM) or provide quick boost response in transient (load step-in at low engine speed).1 Expected by-pass system functionalities for modern gasoline engines The primary functions that a modern gasoline engine by-pass system has to full-fill are the following: .Provide a smooth and progressive opening of the valve in order to avoid abrupt air – and therefore torque – fluctuations that may be perceived by the driver as the by-pass system needs to start opening to control boost pressure (typically during the second phase of low end full load transients). or take-off after idling. but also about the velocity and direction of the waste-gated flow and the likelihood to initiate fast catalyst light off. design parameters and the different constraints and targets are given. engine exhaust pressure and temperature. These examples illustrate the interest of a newly developed valve solution taking the form of a monoblock spheroid arm and valve (1 piece in opposition to widely used 2 piece arm and valve construction). The goal here is to enable so called “de-throttling” strategies which in turn reduce the pumping losses of the engine.. 139 . . . Following complementary functions also have to be taken into account while designing the by-pass system: . or sudden drop of boost demand at high engine speeds need a fast response of the waste-gate valve to provide a prompt feed-back to the driver while keeping emissions under control. . 2. .Maintain sealing in closed valve operation in order to maximise gas energy directed to turbine wheel. The sizing allows the proper control of system critical parameters but also depends on the matching of the turbocharger. engine inlet pressure and temperature. thus improving slightly its efficiency and fuel economy.Example of the typical simulations methodologies used to study the interactions between valve concept. turbocharger speed. This requirement is not only about the mass flow by-passed directly to the catalyst. The closed valve sealing is also important to entertain high instantaneous exhaust pressure peaks which are key to recover max possible energy form exhaust gases into turbocharger power.).Promote the minimum heat transfer to the surroundings and hence the maximum exhaust gas heat / temperature transport directly to catalyst in the case of cold start as measured during typical emission cycle.Allow an easily controllable quasi steady state part opening of the valve to enable the turbocharger waste-gate valve to partially control the boost pressure level in part load. This requirement basically results in the sizing of the waste-gate port (in terms of mass flow Vs pressure delta characteristic).Ensure full control of over all environmental and operating ranges of the critical physical parameters that must be maintained in order to protect vehicle/engine/turbocharger integrity (Engine maximum mean effective pressure. Events like stop and start followed by a quick driver torque demand. The key customer impacted criteria is therefore here low end transient and steady state performance. The key customer characteristic impacted here is therefore vehicle fuel economy. Prevent noise. The by-pass can be seen as a way to vary the exhaust gas flow to the catalyst so that it is favourable to light off in typical cold start conditions and in different way in highly loaded conditions to rather lower the direct exposure of the catalyst to the waste-gated flow. The industry trend is clearly to reduce displacement and to have a lower cylinder count while increasing engine output in terms of specific torque and power. mass and cost. many low engine speed / low engine load operating points are typically operated with a part open waste-gate. This requirement. which is heavily influenced by turbine housing and packaging design constraints. using a downsized engine in a comparable vehicle results in simply having a higher average operating load. . As engine operating conditions typically happen while vehicle is moving at very low speed or even standing still. As a consequence of the control strategies OEMs are putting in place to meet fuel economy and fast response requirements.Resist harsh engine boundary conditions.Limit turbocharger outlet temperatures on highly loaded point once the catalyst has reached steady state temperature. can adversely affect the boundary conditions at the wheel outlet. As by-pass system typically has to connect turbine inlet and outlet. which is a priori contradictory with the necessity to promote quick catalyst light-off. . any noise from the kinematic chain has to be prevented. packaging. . With the high pressure fluctuations and engine vibration encountered on modern and downsized engines. aims to limit the stress put into the catalyst under sustained durability. Finally. This results in a significant increase of engine exhaust pressure peaks which have more than doubled in the last 5 to 6 years. By pass function Figure 1: Key by-pass functionalities 140 . thus increasing the time spent at higher pressure. temperature and vibration. . design of the waste-gate port and nature of the flow in partial or full opening of the valve usually results in wheel outlet perturbations.. This drives a trend of increasing engine vibration as well as maximum exhaust gas temperatures.Promote usage of right sized actuators to limit their impact on current consumption. This phenomenon. any kinematic joint operating with significant clearance is likely to generate intermittent contact and shocks likely to emit air- borne noises or to trigger structure borne noises. therefore impact energy recovered by the wheel.Prevent turbocharger wheel outlet flow perturbation. self heating and its impact on sustainable maximum under hood temperatures. Emission and fuel economy standards are also driving a clear trend on modern gasoline engine which is mostly known as downsizing. A final key environmental factor for valve design is the exhaust peak temperature as well as the maximum positive and negative gradient the engine may generate. They target the driveability of naturally aspirated superior displacement and cylinder count engines like 4 cylinders 1. Note this maximum exhaust temperature selection is also linked to high power combustion strategies which in turn can also affect exhaust gas oxidation potential.8l to 2. They 141 .3 Using computer fluid dynamic simulations to understand valve loading and inherent flow evolution Allowing an easily controllable by-pass valve in near closed position under the boundary conditions described in the previous chapter remains a challenge with the traditional flat poppet design. Note this effect is actually desired since it allows using higher efficiency operating points. As a final effect on valve pressure boundaries.2 Typical boundary conditions for valve operation In order to design turbocharger by-pass valves for gasoline engine application. In order to understand more closely the behaviour of typical waste-gate valves in these conditions. but still facing difficulties and instabilities. OEMs have been evolving from basic pressure actuated valves to vacuum and even electrically actuated valves in order to have more authority and control on such valves. As explained earlier.5l engines are for example being developed and released by OEMS. The key environmental boundaries resides in the description of pressures. This results in driving a higher proportion of maximum engine output to run typical steady state points. In order to keep the same low end full load torque.0l in the case of the above example. exhaust gas pressures have evolved from typical 200 kPa average and 350 kPa peaks to 380 kPa average and 500 or even 580 kPa peaks. Finally. cases applying 1020°C or even higher temperatures are getting released as well. The lower cylinder count also results mechanically in higher exhaust peak pressures to reach the same mean effective pressure. These engines feature less cylinder count and less displacement.2. those engines are typically used in vehicle masses which were previously served with bigger engines. fuel economy regulations are promoting the use of downsized engines. The peaks are also changing in frequency because of the cylinder count and the gearbox matching strategies. The undesired effect is that the overall duty cycle of the engine and the turbocharger get increased to operate more and more time in medium to high loaded points. The acceleration levels are typically increasing and covering lower firing frequencies. As such many new 3 cylinders 1 to 1. these engines are mechanically resulting in higher maximum cylinder mean effective pressures (above 20b BMEP). Another factor affecting valve operation directly linked to engine superior loading and lower cylinder count is the vibration. Note the trends are less generic on vibrations than they are on exhaust pulsations as engine architecture (balancing shafts or flywheel) is largely affecting them. 2. While exhaust temperatures in the 980°C remain main stream. computer fluid dynamic simulations have been developed. temperatures and vibrations. This oxidation potential has to be taken into account with regard to the adverse effect it may have together with mass flow and temperature on by-pass mechanism functionality over time. There is no general trend again here other than the trade-off between system cost and high power fuel efficiency OEMs are subject to. especially at low openings. These flow and load evolutions near the fully closed positions definitely promote valve instabilities and generate extreme difficulties to control the turbocharger. valve opening and operating conditions. Figure 2 below illustrates the static pressure contour of a flat poppet as a function of opening. the new proposed valve features a conical seat and a toroidal valve shape interface. 142 . Secondly. While a traditional poppet relies on a flat surface contact and a significant clearance between the poppet and the arm to accommodate for tolerances and thermal relative displacements. and then reaches a maximum before decreasing again. As can be seen a flat valve results in two consequences that are detrimental to controllability. as suggested in the illustrations of Figure 3. The simulation also allows better understanding what governs the valve loading behaviour. This interface allows reducing the area evolution which is subject to turbine inlet pressure when the valve opens. the valve loading also sees an abrupt increase when the valve starts to open. Figure 2: Flat poppet flow & loading as a function of valve angle (top 2 graphs) and valve static pressure contours (CFD contour plots) Based on this understanding. which is another key factor to look at to assess valve controllability. the valve flow increases very aggressively with the valve angle. a new concept of valve has been engineered in order to provide a smoother evolution of flow and load as a function of valve opening.allow formulating the nature of the waste-gated flow as a function of time. First of all. The key feature of this new valve resides in a different sealing interface between the valve and the seat. The overall effect is to decrease the load by approximately 30% compared to the original valve. not only is the valve angle required to achieve the same flow higher. but the load required to keep the valve in this position is also lower.Another interest of this interface is that it naturally results in a smoother geometric area increase while the valve is opening and provides a “self centering” ability which no longer requires clearance between the poppet and the arm. Finally the valve features a spheroid protrusion (Patent pending) inside the waste-gate channel that contributes to splitting the flow around 360 degrees. This is a clear advantage for valve and kinematic chain durability. Note that alternative protrusion shapes have also been investigated (conical as well as a complex non axi-symetrical 3D design). which is definitely helping the ability to maintain the valve in near closed position whether it is with vacuum or electric actuator. 143 . These shapes still provide the nice features of monotonic load evolution as well as a better flow progressivity than the baseline flat poppet. a nice monotonic load profile is achieved. the best achieved characteristic is definitely achieved with the spheroid protrusion version. Figure 3: Valve sealing interface and impact on load evolution Figure 4 presents the results in terms of flow and load that are achieved with this new concept. confirming the results anticipated by the CFD. As a consequence the spheroid protrusion has been selected to proceed to further on engine and vehicle controllability assessment. As can be seen. It also decreases drastically the impact pressure pulsations may have on the valve position or load fluctuation. Finally. The resulting proposed conical seat to toroidal poppet monoblock arm and valves have also been tested on a cold flow bench. However. The combined effect of the monoblock arm and valve result in a smoother effective area progressivity. Note that this test has been performed using the target pneumatic vacuum actuator in order to directly understand the relationship that exists between the commanded vacuum and the resulting waste-gate effective area. Figure 5: Experimental results on waste-gate effective area (cold bench) 144 . This metric basically combines the effects of flow and load progressivity. the monoblock arm and valve finally catches up the full open effective area on the last 20 mb. On the other side of the graph towards the full open position. This is due to the fact that the valve spheroid protrusion finally gets totally out of the channel in the last degrees of opening of the valve. Figure 4: Flat poppet Vs Monoblock valve flow and load evolution (CFD) Figure 5 presents the experimental results that have been collected on the cold flow bench. since vacuum actually sets a force on the actuator and not a position. in order to ensure full functionality.4 Using combined full stage finite element analysis and multi-body simulation to ensure valve functionality in critical thermal transient conditions As explained earlier. Note this multi-body simulation also uses as an input the valve loading computed through CFD and explained previously. FEA and multi-body simulation 145 . Figure 6: Combining CFD.2. The study then consists in retrieving a worst case deformation of all the parts and their relative position at the critical time step of the most severe transient thermal event. The figure 6 explains the simulation flow logic that is used to assess the self centering of the valve in hot conditions. the proposed monoblock toroidal portion of the arm and valve has to automatically centre in the turbine housing conical seat. This critical hot deformed shape is then imported into a multi-body simulation tool in order to assess if the valve does naturally self centre in the seat while it is being closed through actuator force. As the valve actually operates rather in a hot temperature environment. Those key geometries include seat and valve deformation in the region of sealed contact. The goal of the finite element study is to understand what are the thermal transient events and critical time through those events that impact the deformation of critical geometries entering into valve self centering stack-up. To that end. effects of thermal distortions and friction coefficients while the valve approaches the seat also had to be taken into account. Complex 3 dimensional stack-up calculations have been performed in the cold state to ensure clearances in the kinematic chain allow the valve to self centre without any hyper static conflict. but also the relative displacement between the arm and valve guiding and the seat. finite element analysis of the turbine stage (including arm & valve) has been performed. especially at the valve arm to bushing contact. Since it removes clearance between the valve and the arm. noise. resulting in no noticeable low end torque drift on the vehicles. This translates in vehicle in an improved driving experience. both on performance and controllability. from performance target setting to kinematics design. 3 SUMMARY AND OUTLOOK In this paper. for 2017. Indeed. the monoblock arm and valve design has been refined in order to achieve an optimum design in terms of performance. durability. Looking at the next evolution of emission regulation. regular valve closed turbine stage permeability measurements have been performed during vehicle endurance tests. This kinematic wear robustness and stroke drift improvement as well as the stable and low valve closed leakage allows to keep the control system closer to the new state over the whole life of the vehicle (same command needed at the actuator for the same engine operating point). durability and performance benefits. the monoblock construction introduction also solved noise issues linked with valve rattling against arm in mid open positions. did enable to achieve a stable control strategy with part open waste gate in part load. and the need to control emissions at higher load with high accuracy will require further enhancement on turbocharging control chains. allowing reducing significantly kinematic wear.As an outcome of these simulations. With breakthrough results. EU6. It has demonstrated a remarkable robustness. Having less actuator stroke drift over engine life not only allows to keep the same actuator position to valve position characteristic. the methodology used to develop the new monoblock spheroid valve design has been presented. Indeed.2 regulation is expected to introduce a new emission cycle (WLTC) with much higher dynamics and load than NEDC cycle. Real Drive Emissions). in addition to the need to respect emission during real drive test (RDE. this design is about to enter production on a 3 cylinder gas application. The new valve brings significant controllability. This results in having a stable low end torque performance as well as a stable dynamic response and driving comfort. therefore enabling the customer to meet its driveability and fuel economy targets. but also ensures a constant available force to keep the valve closed. This results in an actuator stroke drift (difference between the actuator position in closed valve position at new and after test) which is limited to less than 0. They have demonstrated very low and stable valve leakage behaviour over more than 1000h on multiple units. noise and controllability. Finally the controllability enhancement provided by the smoother flow progressivity and the monotonic evolution of valve loading as a function of valve opening. Overall performance benefit versus previous generation flat poppet has been exposed.2 mm over more than 750h of engine test on dynamometer bench where a traditional arm and valve used to result in up to 2 mm stroke drift due to kinematic wear. challenges for powertrain designers seem even more emphasized. less sensitivity to noise while enabling world class fuel consumption and emissions meetings EU6 targets. 146 . This higher dynamics. NOMENCLATURE BB ball bearing BMEP brake mean effective pressure BSFC brake specific fuel consumption . In this study. P Dowell Ford Motor Company. 2014 149 . France 1. ABSTRACT Ball bearing turbocharger technology has started to be adopted for mass-production engines due to the potential benefit in transient performance and fuel consumption. S Akehurst. and that further engine calibration is a must before any fair evaluation of the BSFC benefit can be done. the benefits of using a ball bearing turbocharger compared to a conventional journal bearing turbocharger were identified first in simulation and then validated in a back to back comparison of two otherwise identical turbochargers through extensive experimental analysis. UK G Capon. T Duda. The fuel consumption of the engine was greatly reduced by the ball bearing turbocharger. the low friction of the ball bearing allows the turbocharger to accelerate faster so that the engine can be supplied with boost pressure more quickly following a transient torque request and under steady state offers reduced engine back pressure. UK P Davies Honeywell Turbo Technologies. However. specific heat capacity at turbine condition EGR exhaust gas recirculation axial thrust force JB journal bearing journal bearing correction factor thrust bearing correction factor _______________________________________ © The author(s) and/or their employer(s).Experimental and analytical investigation of implementing a ball bearing turbocharger on a production diesel engine Q Zhang. specific heat capacity at compressor condition . but the full potential of the ball bearing turbocharger in terms of the transient performance can be further exploited by recalibrating the engine. C Brace University of Bath. R Burke. closer scrutiny reveals that the insufficient EGR rate of the ball bearing turbocharger equipped engine was the main cause of the reduced BSFC. which can reduce engine fuel consumption. The hot engine transient response was also improved. The cold start engine performance was significantly improved as the ball bearing turbocharger was able to boost the engine to the full load level within a few engine cycles. Compared to the conventional journal bearing. oil supply delay and hot shutdown. it must be tolerant of high thrust loading. oil contaminants. turbine efficiency on gas stand map turbocharger mechanical efficiency . the technology has become mature enough to be supplied to the mainstream market. On the other hand. and the high fuel price can well justify the cost for the technologies providing better fuel efficiency. isentropic turbine enthalpy change . One of such technologies is the implementation of the ball bearing in the turbocharging system. the turbocharging technology is becoming the standard component on most Diesel engines and many gasoline engines. The conventional journal bearing solution has been able to fulfil these requirements and is the prevalent solution nowadays. . compressor mass flow rate turbocharger oil mass flow turbine mass flow rate NEDC new European drive cycle turbocharger speed . After years of consistent development. the ball bearing system has significantly lower friction and offers better fuel economy and faster transient response. INTRODUCTION As a cost effective method to increase the power density of internal combustion engines. turbocharger inlet/outlet pressures turbine inlet pressure turbine outlet pressure compressor inlet temperature compressor outlet temperature turbine inlet temperature VGT variable geometry turbine WLTC worldwide harmonized light vehicles test cycle compressor power power of turbocharger mechanical loss turbine power heat capacity ratio ΔH compressor enthalpy change ΔH . However. large commercial vehicles and racing cars. A typical turbocharger Figure 1 Turbocharger Ball Bearing ball bearing cartridge structure is Cartridge w/o outer ring shown in Figure 1. The bearing system of a turbocharger is a crucial component as it represents an energy loss in the transmission of power from the turbine to the compressor. the high precision requirement and the high cost has limited its application to niche products. The high driver expectation and strict emission legislation have never ceased to push this technology even further. Many new technologies are becoming available and affordable because of the intensive research input and the surging fuel price. . At the same time. turbine efficiency w/o mechanical loss oil viscosity 2. (Source: Schaeffler AG (1)) 150 . full load and transient conditions.(3) which depends on the internal geometries of the journal and thrust bearings and the oil viscosity. The transient response.2L Diesel engine in a Ricardo Wave simulation environment and on a transient dynamometer. 1− − − − − − (See nomenclature for more) For the purpose of simulating a particular turbocharger on a particular engine. = = ΔH . = 2 The mechanical efficiency of the bearing system is determined by the ratio of compressor work to turbine work = = 3 + Where is defined by the numerator of equation 1 and ℎ is determined through the model developed by Serrano et al. in the context of this work which is focused on the benefits from a particular bearing system. This model is 151 . Fuel consumption should also see a moderate reduction due to the reduced engine exhaust manifold pressure (2). 1 . this is sufficient as the requirements are to estimate the overall transfer of power from the turbine to the compressor. The aim of this study was to provide a back to back comparison of journal and ball bearing technology both experimentally and in simulation. especially of the cold engine. A series of experiments were then conducted covering part load. However. This ensures much lower friction performance in a wide speed and temperature range. The ball bearing can also take the thrust load. .Compared to the conventional journal bearing cartridge. as a consequence the turbine map includes the mechanical efficiency of the bearing system. Two aerodynamically identical turbochargers were implemented on a 2. It is determined in this way to avoid the effects of heat transfer in the turbine and is based on the assumption of adiabatic operation of the compressor which is justified at high operating speeds. ΔH . is expected to be greatly improved. − . BALL BEARING MODEL The characteristic efficiency map of the turbocharger turbine is commonly determined on a gas stand facility based on the measured enthalpy rise in the compressor (equation 1). the turbocharger shaft is supported by the two ball bearing assemblies in place of floating metal bushes. However. 3. this is achieved through equation 2 . eliminating the need for a separate thrust bearing which can further reduce the friction. the turbine map needs to be corrected to represent only the aerodynamic performance. summarised in equation 4. = + 4 : : : ℎ : ℎ Previous studies of ball bearing friction (4. Therefore. 5) suggest friction reductions of around 50% compared to journal bearings. It is expected that the error introduced by the friction model will only account for a small fraction of the errors by a 1-D model and that the simulation should. The controller will adjust the VGT rack position and the EGR vale in order to meet these two targets.5:1 Max Speed [rpm] 4900 Max Torque [Nm] 385 Max Cylinder Pressure [bar] 160 Max Turbocharger Speed [krpm] 213 Max Pre-Turbine Temperature [°C] 830 Max. The model has been applied to the manufacturer supplied map using the conditions of oil temperature and viscosity rating under which the map was originally measured on gas stand. In this way a turbine efficiency map separated from bearing friction has been calculated. . Compressor Outlet Temperature [°C] 180 152 . the only difference between builds was the turbocharger bearing and the engine calibration remained unchanged. due to the lack of measured bearing friction data for the specific turbochargers used in this study. Turbocharged. . Table 1 Test Engine and Turbocharger Specifications Engine Configuration L4. EXPERIMENTS A 2. 4.2 Bore [mm] 86 Stroke [mm] 94. but for full details the reader is directed to the original publication. This is important in that the engine controller has two key set points of inlet manifold pressure and intake air mass flow rate. the friction model introduced here will be applied to the 1-D gas dynamic model. . or are to be examined in the next phase of the study. to some extent. The specifications of the selected engine and turbochargers are summarized in Table 1. allow us to look at different control logics that are either difficult to check experimentally. In both cases. Intercooled Displacement [L] 2.6 Connecting Rod Length [mm] 155 Compression Ratio 15.2L Diesel engine was chosen as the test engine and the default turbocharger and an updated device with ball bearing cartridge were used in turn on the engine. . Hot engine torque transients . High frequency pressure measurements were taken at crucial locations such as the exhaust ports.Same test as the cold start torque transients at hot engine condition. post/pre turbocharger so that hot air pressure rise after transients and exhaust gas pulsation details can be recorded and analysed. Table 2 List of Types of Sensors Pressure PTX: Druck and RS Temperature K Type Thermocouple Mass Flow ABB Sensyflow Fuel Flow (steady state) Gravimetric Fuel Balance Turbocharger speed MicroEpsillon Eddy Current Hot Gas Pressure Kistler 4049 Cold Air Pressure Kistler 4007 The experiments were designed to cover the most pertinent engine operative conditions. The table 2 summarises a list of types of sensors used in this study.The whole air-engine-gas path is monitored by paired temperature and pressure sensors at 40~80Hz. including: Limiting torque curve .The engine was tested for steady state part-load at hot engine condition. Cold start torque transients .The engine was started from 15 °C and was controlled to enter into the full load transient schedule within 20s (time for the fuel beaker to settle). Figure 2 Combined WLTC/NEDC Minimap Points 153 . Part-load steady state points generalized from combined WLTC/NEDC drive cycles (Figure 2).The engine was tested for steady state full load torque at hot engine condition. The journal bearing did not achieve its full load torque within the transient test period as it will only happen when the engine oil is fully warmed up. Therefore.2. The BSFC is consequently also only marginally improved (Figure 3).2.1. RESULTS AND DISCUSSION 5. only representative test results of each of the four tests are presented in the paper. Figure 3 Limiting Torque Comparison and BSFC Reduction (Positive percentage -> Improvement) 5. the air mass flow rates are similar and the full load torque was improved up to 6 Nm at 2000 rpm.3s faster than the journal bearing turbocharger. etc). Transient response 5. mostly from reduced pumping work because the engine back pressure is reduced by the ball bearing turbocharger. Limiting torque curve At full load. exhaust manifold temperature. This is largely because the engine full load condition was often limited by factors other than the turbocharger performance (cylinder pressure. 154 . the ball bearing achieved 90% torque 1. whereas the journal bearing achieved only 75% and only reached 90% in the 6th transient.1. The result in torque performance was significant: the ball bearing turbocharger achieved 90% of the full load torque from the first tip-in. It should also be noted that the full load torque is the same for both turbochargers at this speed and the ball bearing turbocharged engine managed to achieve this torque from the 5th transient. Due to the limit in the paper length. with the same calibrated boost target.The part-load steady state points were selected with a weighting process from the WLTC/NEDC drive cycle simulations using a representative vehicle. Cold torque transient The cold start torque transient tests showed a clear advantage of the ball bearing turbocharger (Figure 4). When looking at the first transient response only. which stabilized at torque of 40 Nm lower. 5. the ball bearing turbocharger does not give an apparent improvement to the engine torque performance. The curves illustrate that the ball bearing turbocharger started from Turbocharger speed 50% higher than the journal bearing turbocharger and maintains a speed at least 10% faster throughout the transient test. Therefore. and the ball bearing stabilized transiently at torque 14 Nm higher. Hot torque transient For a fully warmed up engine. At such conditions.2.2. Figure 5 Hot Engine Torque Transient at 1000 rpm and the Exhaust Manifold Pressure In both cases. the ball bearing turbocharger generates a lower engine back pressure. However.2s (41% reduction). the ball 155 . The transient response comparison at 1000 rpm is shown in Figure 5. The time to 90% JB torque is reduced by 1. This is a calibration setting for the engine to generate high EGR gas at low engine load and/or to prepare a fast spinning turbocharger for the transients. the difference in shaft friction becomes smaller and the friction loss counts as a smaller proportion of the total work done by the turbine. Although with lower back pressure. the ball bearing turbocharger still provides better torque transient at low engine speed. especially at higher speed. It is clear that the ball bearing turbocharged engine responds faster than the journal bearing turbocharged engine. the EGR valve in the ball bearing turbocharged engine was more open to ensure that similar EGR rate can be achieved with the lower back pressure (EGR rate 44% compared to 42% for JB). Figure 4 Cold Start Torque Transient at 1250 rpm and the Turbospeed Difference 5. the VGT vanes are at the fully closed position before the transient. The transient response comparison at 1500 rpm is shown in Figure 6. the torque rises are very similar for ball bearing and journal bearing (even slightly faster with the journal bearing). and both reached the same stabilised torque (BMEP 21.8 krpm compared to 26 krpm): this translates to faster torque rise. Figure 6 Hot Engine Torque Transient at 1500 rpm and Relevant Parameters 156 .bearing turbocharger has a higher speed prior to the transient (35. Unlike at 1000rpm.5 bar). The back pressure is reduced across the load range by 15 . the journal bearing turbocharger has faster exhaust pressure build up to accelerate the turbocharger and therefore to provide boost. The EGR valve is more opened in the BB case such as to meet the EGR requirements (61% open compared to 42% for JB). However. However. the exhaust energy is more abundant so that the more closed VGT vane position leads to a higher turbocharger speed and back pressure for the JB turbocharger.Before the transient. ball bearing turbocharger is the better option. These differences are between 3-11% improvement through the turbocharger and it is clear that this benefit is not solely a result of reduced turbocharger friction. but due to interactions with other engine systems.29 kPa. the benefit of higher back pressure with the journal bearing turbocharger balances the benefit of lower shaft friction and lower pumping work of the ball bearing turbocharger. 5. in terms of exhaust manifold component durability. Compared to the 1000 rpm transient test. When the transient happens. when the test results reveal that the ball bearing turbocharger gives a large fuel consumption benefit. fuel consumption and controllability. Figure 7 BSFC Comparison of Minimap Points at 1500 rpm (Positive percentage -> Improvement) Figure 7 shows the raw fuel consumption benefits for the range of engine loads at 1500rpm resulting from the back to back comparison of the two turbocharger bearing technologies. The overshoot in turbocharger speed/boost also leads to an obvious overshoot in torque in the journal bearing transient test. it is very difficult to generate a convincing engine fuel consumption performance in the drive cycle part- load test. it is clear that without re-calibrating the engine for the ball bearing turbocharger. the ball bearing turbocharger was able to generate sufficient pressure in the inlet manifold to meet the calibration target (VGT 58% shut compared to 78% for JB). In terms of the torque rise. Part-load points fuel consumption The part-load quasi steady drive cycle tests can be seen as a crucial indicator of the engine fuel consumption. Figure 8 shows the impact of the turbocharger bearing on back pressure and EGR flow. This reduction in 157 .3. Figure 8 Back Pressure and EGR Rate Difference between JB and BB Equipped Engine at 1500 rpm According to the research on the diesel engine back pressure (6) and EGR rate (7). A crude calculation of the acquired BSFC benefit through back pressure and insufficient EGR flow is illustrated in Figure 9. there will be a BSFC increase of around 0. for every 1 Bar of increased back pressure. notably between 50-250Nm where the largest fuel consumption gains are made.backpressure has a significant effect on EGR rate. Figure 9 BSFC Reduction Analysis Compared to Measurement 158 . the BSFC penalty is around 50 g/kWh. while for 1% of EGR rate.3%. but also NOx and soot emissions. This reduction in EGR rate of up to 15percentage points would have a large impact on fuel economy. The dilemma suggested yet an optimal solution: by choosing a slightly smaller turbo-machinery. Ideally the engine should be at least recalibrated for the newly fit ball bearing turbocharger. Therefore. Figure 10 is a compiled graph of the rest of the part load points at 1000 rpm. 2500 rpm and 3000 rpm Although a fair comparison was intended between the two turbochargers. 2000 rpm. However such a study is beyond the scope of this paper. However. 2000 rpm. this process was not permitted in the time span of this project. 2500 rpm and 3000 rpm. It should be noted that such BSFC reduction is of limited practical use as the engine will still be required to meet NOx limits which on the current engine will require similar EGR rates as the JB configuration. The transient response can be further improved in the meantime. it was decided that the engine calibration remained unchanged for both turbochargers. To meet such requirements would involve the complete re-optimisation of the engine controller and notably changing the targets for boost pressure and EGR flow based on the emissions/fuel economy trade off. This analysis allows a crude isolation of the fuel economy benefits due to higher efficiency of the turbocharger and reduction of the EGR flow. with a prediction of 5-15g/kWh benefit in fuel consumption from the reduction in back pressure. This gives a comparatively conservative demonstration of 159 . it was later realized that a perfect back to back comparison is near to impossible in the case of the engine-turbocharger system: any parameter changed would have a chain effect on the parameters in the engine-turbocharger loop. . The calculated BSFC reduction correlates reasonably well with the measured BSFC reduction (Total and Measured BSFC Difference). Figure 10 BSFC Comparison of Minimap Points at 1000 rpm. the higher back pressure is expected to be able to drive the EGR gas back to the inlet manifold which is not so highly boosted. This is consistent with the results shown at 1500rpm. 6. The model was run in conjunction with Mathworks MATLAB Simulink to allow a high level implementation of the engine controller software and was calibrated against both high and part load engine operating conditions running up to 3000 rpm. As a result. Hot torque transient simulation The transient simulation involves improving the 1500 rpm ball bearing turbocharger torque response. and turbine map was adjusted using the Valencia model to represent the ball bearing turbocharger. a simulation study was conducted in an engine wave action model of the 2. It was observed that the ball bearing turbocharger produced slower torque rise at higher speed compared to the journal bearing due to the fact that the more efficient ball bearing turbocharger rests in lower turbocharger speed compared to the journal bearing turbocharger. the ball bearing turbocharger generated a similar level of engine back pressure which implied a similar turbocharger speed. Therefore. so that the VGT will be at a more closed position to prepare for the torque tip-in. The simulation result is illustrated in the Figure 11.2L Diesel engine using Ricardo WAVE software.the benefit of the ball bearing turbocharger as the calibration was optimized only for the original journal bearing turbocharger and the simulation study was designed to cover some of the weakness of the experimental study. 6. This suggests that an engine re-calibration would benefit the ball bearing turbocharged engine in the transient performance in a wider speed range. Figure 11 Simulated Hot Engine Torque Transient at 1500 rpm and the Exhaust Manifold Pressure As shown in the figure. supplied by the turbocharger manufacturer. with a higher boost target. The journal bearing turbocharger was simulated by characteristic maps of the compressor and turbine.1. the boost target of the ball bearing turbocharged engine is adjusted to a higher value. the torque generated by the ball bearing turbocharged engine has the fastest climb compared to the bench mark engine and the ball bearing turbocharger running the original engine calibration. SIMULATION STUDY Since the time and calibration effort required to perform a perfect back to back comparison of the ball bearing and journal bearing device are not allowed in this first phase of the project. 160 . 2. To ensure an accurate as possible prediction of this steady state fuel consumption simulation study. the simulation was designed to achieve the EGR rate targets instead of the mass flow targets and the results are supposed to reveal the true fuel consumption benefits can be achieved by implementing a ball bearing turbocharger on the assumption that similar EGR rate level would produce similar level of NOx emission. 161 . Therefore. 7. the original journal bearing turbocharged engine model was especially calibrated manually for the FMEP of each selected brake torque. the engine will not achieve the EGR rate required for emission control.6. the average fuel consumption benefit of the ball bearing turbocharger settled to a reasonable value of 2. Part-load points simulations The part-load points simulation focuses on one of the problematic speed that produced overestimations in fuel consumption benefits. Figure 12 Simulated BSFC Comparison of part-load points at 1500 rpm (Positive percentage -> Improvement) The simulation results demonstrated that with a similar level of EGR rate. CONCLUSIONS In this paper. The results are shown as in the Figure 12 below. An aerodynamically identical journal bearing turbocharger was also tested as the benchmark. The results are also in line with the theoretical fuel consumption benefit from lowered engine back pressure. It was pointed out in the analysis in section 5.3 that when running the ball bearing combined with original engine calibration.5% BSFC reduction (without the 250Nm negative result taken into account). a novel turbocharger equipped with ball bearing rotor was installed on a production engine to evaluate the benefit in terms of fuel economy and engine transient response. C. E. D. Sadeghi. Michel. Griffith. P. “Theoretical and experimental study of mechanical losses in automotive turbochargers”.The test and simulation results showed that 1. 2013 (4) B. Ball Bearing. Donaldson. 2013 (2) Honeywell Turbo Technologies. Large fuel consumption benefits can be seen when running part-load steady state tests because of the interactions with EGR system. and A. E. Appliance of high EGR rates with a short and long route EGR system on a Heavy Duty diesel engine. P. 2013 (6) P.. Australia (7) M. Ball Bearing goes Mainstream. Defence Science and Technology Organisation. Serrano. 3. C. The Effect of Back Pressure on the Operation of a Diesel Engine.honeywell. IQPC. J. Davies. Marsal. SAE Paper Number 2007-01-4235. 2011. 2007 162 . REFERENCE LIST (1) P. Garcia-Cuevas. Tiseira.. Slaughter. Archer and J. Lefebvre. Simulation results showed that with small modifications to the engine control strategy. 4. 2007 (5) M.. vol 135/041102-1. fuel consumption benefit of 2. pp. An engine control system re-optimisation is needed to make a true back- to-back comparison.D. F. Germany.5% could be gained. Barthelet. R. Brouwer. Wiesbaden. R. and Mavrosakis. S. [online] Available at: http://turbo. et al. vol. Olmeda. Maritime Platforms Division. Lancaster. Hield. van Aken. Whirl and Friction Characteristics of High Speed Floating Ring and Ball Bearing Turbochargers. 55. A. Transactions of the ASME: Journal of Tribology. There is a significant benefit of cold start transient response can be gained by implementing the ball bearing turbocharger. SAE Paper Number 2007-01-0906. 2.com/our-technologies/ball-bearing/ (Last accessed: Oct 21st 2013) (3) J. R. Energy. 888-898. “Applying Ball Bearings to the Series Turbochargers for the Caterpillar® Heavy-Duty On- Highway Truck Engines”. L. P. M. 2014 163 . 1 INTRODUCTION Downsizing the internal combustion engine can reduce the fuel consumption by moving the fuel efficient zones closer to road driving conditions. A mass air flow ratio of over 80:1 from rated power to idle is typical for gasoline engines. 5]. the VGT turbocharger can allow all the exhaust gas to pass through __________________ © Huayin Tang. The time required to reach 50% of maximum torque rise (T50) was improved by up to 0. a model based control strategy may be required. and reach the required engine torque [4. the transient response of conventional turbocharged gasoline engines is usually slow compared to naturally aspired engines. reducing the available brake torque during the early part of the transient. Fully closing the VGT resulted in high exhaust back pressure and low volumetric efficiency. It highlights the demand for varying the characteristics of the boosting systems on gasoline engines [3].5%) whilst the turbocharger acceleration was maintained. In addition. Therefore.Optimisation of transient response of a gasoline engine with variable geometry turbine turbocharger H Tang. In addition. achieve the target boost pressure. This suggests that a simple boost pressure feedback control will likely not deliver optimised performance due to the excessive exhaust back pressure. This is due to the period of time required to accelerate the turbocharger. UK S Garrett Cummins Turbo Technologies Limited. This can be achieved by turbocharging [1]. S Akehurst. In order to explore the transient operation without any limitation imposed by the production control strategy. C J Brace University of Bath. The trade-off between the responses in different stages in the transient event has been illustrated. the VGT can change the gas velocity and flow angle to vary the turbine characteristics [6]. One of the available boosting systems that have the potential to achieve rapid transient response and high fuel efficiency over a wide flow range is the variable geometry turbine (VGT) turbocharger [3]. Variable Geometry Turbine (VGT) turbochargers offer a route to improve the transient response. Compared to fixed geometry turbine. UK L Smith Jaguar Land Rover Limited.54s (35. an on line search was conducted using a series of open loop actuator trajectories applied to a VGT turbocharger installed on a gasoline engine. in contrast to the ratio of approximately 6:1 for passenger car Diesel engines [2]. UK ABSTRACT Maintaining transient torque response is challenging on turbocharged engines because of the period of time required to accelerate the turbocharger. The blow-off valve on the compressor housing was deactivated. the intake and exhaust valve timings were kept at the maximum overlap positions in the transient test. 164 . turbocharged gasoline engine with variable intake and exhaust-valve timing system. a series of open loop VGT actuator trajectories generated in dSpace environment were provided to an electric actuator using CAN message to control the VGT position. etc. the application of VGT turbocharger on gasoline engines is more challenging. The throttle was fully opened at the start of the transient test. Slow measurements for pressures. the specific enthalpy gradient required to drive the compressor can be reduced [7].5 degree crank angle resolution. flow. temperatures. The variable geometry nozzle turbine (VNT) [12]. and the original production level spark timing control strategy was used.1 second.the turbine. several types of VGT are available. Before the start of each transient test. respectively. The VNT type offers higher peak efficiency and larger flow range. As a result. As a result. The coolant temperature at engine outlet and the air temperature at intercooler outlet were maintained at approximately 90˚C and 30˚C. were recorded at 0. turbocharger speed as well as pressures and temperatures at turbine and compressor inlet and outlet. injection pressures. both steady state and transient control strategy can be improved due to the higher control flexibility [9].1 Test set-up The experiments were performed on a 2. This is because the exhaust gas temperature of gasoline engines can reach over 1000°C. Moreover. 11]. Fast measurements of cylinder pressures. the VGT concept is applied predominantly on the Diesel engine. the VGT can achieve higher turbine efficiency over a wider operating range [8]. the transient response of a gasoline engine fitted with a VNT turbocharger has been optimised in this study. and variable flow turbine (VFT) [14] can vary the effective flow area radially. 2 METHODOLOGY 2. the engine was settled at 2 bar BMEP for five minutes. due to the complexity of the mechanism.05 – 0. The maximum power output of the tested engine is 150 kW. over 200°C higher than that of Diesel engines [10. The Lambda target was maintained at 1. while the sliding wall with variable axial width [15]. To illustrate the potential of using VGT on gasoline engine. The engine speed was maintained at 2000 rpm. Hence. variable turbine housing throat area (VAT) [13]. The test was carried out on a dynamic engine test bed in the University of Bath. as the available boosting capability is relatively high and the transient response of the engine is crucial at this engine speed. In order to eliminate the interaction between the turbocharger and the valve timing. The original fixed geometry turbine turbocharger on the engine was replaced with a VNT turbocharger. In order to study the effect on transient engine operation unconstrained by possible deficiencies in the feedback controller. and torque were logged at 80Hz. Despite the difficulties. port pressures and temperatures. four-cylinder. The timing of the transient trajectory in dSpace was triggered by the pedal position voltage signal from test bed control system. and the VGT transient trajectory was also started at the same time. The demanded and actual VGT actuator position signals from the CAN message were recorded in the fast measurement system. However. and the response time of the VGT actuator was approximately 0. direct injection.0L. and twin scroll switching type [16] vary the effective flow area axially. 165 . 75% closed. respectively. these three VGT positions were still selected as options because of the potential of accelerating the turbocharger faster.5 seconds was divided into three 0.2. the limit of the number of tests. although 50% of the maximum torque rise (144 Nm higher than the low load operating condition. 60% closed and 90% closed (fully closed) were tested at 2 bar BMEP 2000 rpm. It was found that with the VGT position between 60% and fully closed after tip-in.5 second stages. which is crucial in the engine transient response. 216 transient tests with different VGT actuator trajectories were carried out. each of which had a number of optional VGT positions. the steady state fuel consumption and turbocharger speed at three VGT positions. Therefore.7 krpm. 85% and 90% closed. 32 Nm) can not be achieved within 1.3 krpm. However.5 seconds with the VGT position 80%. six different VGT positions (from 60% to 90% closed) were selected as options at each one of the three transient stages after tip-in. a range of VGT actuator trajectories with single step change.9 krpm and 19. In order to define the boundaries for the VGT positions in the transient operation.05%. 85% closed and 90% fully closed). it was chosen that the 1.5 second into transient. have been tested. 80% closed. The averaged fuel consumptions at the three operating points were the same considering that the accuracy of the measurements is ±0. shown in figure 1. 1.1 bar boost pressure downstream the compressor can be achieved at 1. the averaged turbocharger speeds at the three VGT positions were 7. 12. and the VGT actuator response time.5 seconds. 16% closed (fully open). Therefore. In addition. 70% closed. 40% closed. the VGT position was chosen to be fully closed at low load operating point before transient trajectory was triggered. Due to the need of having more than two stages in the transient to optimise the dynamic response of the system. 60% closed. The purpose of this experiment was to analyse the transient behaviour of the engine in the first 1.2 VGT actuator trajectories To determine the VGT position before tip-in. Thus. from fully opened to fully closed (16% fully open. Figure 1 Torque responses of the single step change VGT actuator trajectories. 1%.1%). the turbine inlet pressure was approximately 0. respectively. This is the same for other following figures.0 – 0.5 – 1.0 second) are compared in figure 2. The calculated volumetric efficiency was approximately 0.3 RESULTS AND DISCUSSIONS 3. and VGT setting at third stage (1. This resulted in lower turbine total-to-static pressure ratio and the higher engine volumetric efficiency. VGT setting at second stage (0. With the VGT 60% closed in the second stage. VGT position at first stage (0. With 60% closed VGT. The comparison of torque response at second stage is also shown in figure 3.5 s). the maximum difference in torque at the second stage was up to 17. The VGT position described in the figure is defined as follows: the VGT setting before tip-in.1 Response of engine torque The engine torque responses of six tests with different VGT positions at the second stage (0. Figure 2 Torque responses of the six tests with different VGT settings. shown in figure 5. The VGT positions in the first stage and third stage in the six tests were the same.2 bar lower than that with fully closed VGT.2 seconds delayed compared with the turbine pressure ratio due to the distance between the flow meter and intake manifold. respectively. The instantaneous turbine inlet pressures are shown in figure 4.0 – 1.5 – 1. 166 . The torque and turbocharger speed variations at the end of first stage were within ±0.0 s).5 s).8Nm (14. torque increased immediately after the restriction in the exhaust system was released due to the opening of the VGT. The measured torque has been smoothed using 10-points smoothing.6% and ±2. However. 5 – 1. the turbocharger acceleration rate was 8. Therefore. Figure 4 Instantaneous turbine inlet pressures at the second stage of three tests. This was not preferable in terms of drivability. Figure 3 Torque responses at the second stage (0. it shows a clear trade-off between the torque rise and turbocharger acceleration which affects the torque response at the next stages in the transient event. Compared to other trajectories. On the other hand. 167 . The optimised trajectory is also constrained by the drivability requirements.0 s). and approximately 0. In addition. if the fastest turbocharger acceleration was pursued.5 second after tip-in. a torque drop can be observed when the VGT was returned to 75% closed at the third stage. However. Although the turbocharger acceleration was the fastest. The measured torque has been smoothed using 10-points smoothing. the drivability was also not acceptable.2 second at 1. This delay was approximately 0.35 second was required to recover this torque drop. the slow torque rise at the third stage was also a result of the low turbocharger acceleration at the second stage. the torque was dropped by approximately 10 Nm when the VGT was closed.8% lower than that with fully closed VGT. 9478.0 s with VGT 75% closed. Figure 5 Turbine total-to-static pressure ratio and engine volumetric efficiency. respectively. The fittings at other stages with different VGT settings are showing similar trends. A typical fitting at the third stage with the VGT 75% closed is shown in figure 6.9476 and 0. The maximum deviation between the predicted turbocharger speeds using second order fitting was 2. Figure 6 Fitted responses of the turbocharger speed at 1.5 s and turbocharger speed at 1. the engine torque response is largely affected by the acceleration of the turbocharger. for the tests with the same VGT setting.94%. No significant difference between the first order and the second order fitted curves was observed. turbocharger acceleration rate was analysed.5 second) using the averaged turbocharger speed before tip-in as an input. and they are not presented.2 Response of turbocharger speed On turbocharged gasoline engines. The fitted curves at all stages were then used to predict the turbocharger speed at the end of transient test (1. lower than that using 168 . 3. the R2 of the first order fitting and second order fitting are 0. At each of the three stages in the transient tests. first order and second order curves have been fitted to the turbocharger speeds entering the stage and the turbocharger speed at the end of the stage. Therefore. The turbocharger acceleration was also selected because the T90 can not be reached within 1. 0. A second order polynomial curve has been fitted to the mean values of torque and turbocharger speed rises of the six data groups. Therefore. In addition. the repeatability of the test is acceptable.the first order fitting which was 3.5 seconds relatively linearly dependents on the turbocharger speed at the time entering the stage and the VGT position at the stage. only 4 test points in total are outside the 75% probability regions.5 second. It was also found that in all the three stages. The non-linearity of the VGT turbocharger has been illustrated. Each one of the six groups of data represents the test result of 36 different VGT actuator trajectories that had the same VGT position in the first 0. The ellipses shown in the figure represent the 50% probability region of the distribution of the data having the same VGT position at the first stage.5 second is analysed. The engine torque rise was selected because this is the output of the engine.5 second. The elliptical 50% probability regions of torque and turbocharger speed rises in the first 0. Therefore. This illustrates that. although the VGT turbocharger is flexible to reduce the effective area and to accelerate the exhaust 169 . the difference between the predictions from the two fittings was below 0.3 Optimisation of the transient operation To optimise the transient operation. the slopes of the fitted curves were between 0. the response in the first 0. the turbocharger speed benefit gained from having a higher entering speed was smaller than the difference in speed at the start of the stage.5 seconds is largely affected by the turbocharger acceleration. two parameters were chosen as the indicators. although the higher turbocharger speed at the start of a stage led to higher turbocharger speed at the end of that stage.5 seconds.03%.77%. Therefore. Firstly.5 second stage in the first 1. was small enough to optimise the transient operation. and the transient response after 1. Figure 7 Response curve between the turbocharger speed at the end of the first stage and the torque rise in the first stage. Fully closing the VGT was not beneficial to both the torque rise and turbocharger speed rise. 3. Despite the overlap between the groups because the VGT were close to fully closed position. the chosen stage duration. the rise in turbocharger speed at each 0. In addition. However.5 and 1. it has been illustrated that the turbocharger speed at the end of a stage is strongly correlated to the turbocharger speed at the start of the stage.5 second have been plotted in figure 7. A Pareto optimal front was drawn. it was difficult to calculate the turbine efficiency in a transient test due to the response time of the thermocouples and the heat transfer effect on turbocharger. the calibration of the strategy depends on the requirement of the transient operation. It is also illustrated that. Figure 8 Trade-off between the turbocharger speed at 1. which can be either a fast torque rise or fast turbocharger acceleration at the first stage. The turbocharger characteristic maps are not available. the methodology of optimising the turbocharger control strategy is applicable to other applications. fine tuning of VGT controller and model-based control strategy may be required. therefore. On the other hand. overshoot in VGT position feedback control may result in both slow turbocharger acceleration and slow torque response. This may be a result of low turbine efficiency at fully closed VGT position. It is worth noting that the baseline control strategy with the production FGT turbocharger achieved T50 of 0. Therefore.5 seconds are not shown. to enable the use of VGT turbocharger on gasoline engines.5 second and T50 can not be achieved with VGT position 60% closed at the first stage. the engine response in the first 0. As the response in the first 1. and the turbocharger speed rise was significantly deteriorated.51 s and turbocharger speed of 113 krpm at 1. 170 . it can not be compared with the baseline production turbocharger. opening up the VGT position to below 60% closed only gave marginal benefit on the torque rise at the first stage. Therefore. The VGT turbocharger was not matched to the tested engine. the VGT positions at the second and third stage were optimised. It is likely to be a result of the differences in the turbocharger matching. However. The turbocharger speed at 1. It was found that the highest Pareto efficiency can be achieved by having a VGT position between 70% and 85% closed at the first stage. Thus. because of the non-linearity of the VGT turbocharger.5 second is plotted against T50.5 s and T50. optimum turbocharger speed rise in the first 1.gas velocity.5 seconds was concerned. which can not be achieved by the VGT turbocharger with the tested trajectories. although the fastest torque rise in the first 0. However.5 second can not be improved by fully closing the VGT. shown in figure 8.5 seconds.5 second was reached. Secondly. tests in which 50% of the maximum torque rise can not be reached within 1. the turbocharger acceleration was the main target. In the region 2. The transient VGT control strategy can be optimised such that: a. Figure 9 Pareto optimal front of the trade-off between turbocharger speed at 1. However. The fastest T50 was achieved by opening the VGT to 60% closed at the second stage to release the restriction in the exhaust system and to allow the 50% of maximum torque rise to be reached within 1 second.5 second was relatively lower and the T50 was shorter. The highest turbocharger speed at 1. Therefore. The VGT was closed to intermediate positions at the first stage to accelerate the turbocharger. b. the turbocharger speed can not be recovered from the losses due to the opening of the VGT at the second stage. A significant torque drop can be observed on the torque responses of the majority trajectories in region 1. the two trajectories closed the VGT to intermediate positions (80% and 75% closed) at the first stage and opened the VGT to 70% until T50 was reached. The VGT position was kept at intermediate positions to pursue fastest turbocharger acceleration. shown in figure 10. shown in figure 9. the VGT was kept between 80% and 75% closed in most of the trajectories in region 3.5 seconds and T50 has been illustrated. the turbocharger acceleration was deteriorated and the gap can not be recovered. The trajectory “90-85-60-70” was acceptable because the VGT was not closed aggressively. due to the closing of the VGT to accelerate the turbocharger after T50 was reached. c.5 s and T50.The Pareto optimal curve is divided into three regions. and it was then opened mildly to allow both acceptable turbocharger acceleration and torque build up. a clear trade-off between the turbocharger speed at 1.5 second was achieved in this region although the torque rise was slower. No torque drop is observed. The minimum T50 was achieved by opening up the VGT and releasing the restriction in the exhaust system. In region 1 where the turbocharger speed at 1. Therefore. 171 . With all the possible VGT settings at the third stage. In the third region of the Pareto optimal curve. 25 s T50 improvement (33.5 krpm turbocharger - speed at 1.8 krpm -1. - first 0. despite the flexibility of fully closing the VGT and reducing the effective area to accelerate the exhaust gas and increase the kinetic energy.4 Nm -2. which achieved the fastest torque rise in the first 0.80% closed first stage closed VGT position at 60% closed 60% closed 70% closed before 75% .5%) Improvement of -3. Table 1 presents the optimisation of the VGT actuator trajectories targeting different requirements of the transient operation.6%) (-1. It was found that.7 krpm 0.51 s 0.54 s 0. the turbocharger acceleration and torque response can not be benefitted by fully closing the VGT at all the three stages. the highest Pareto efficiency can not be achieved.5 speed at 1.80% 60% closed 85% closed 75% .5 Nm torque rise in the .5%) (16.5 second was crucial.80% second and third before T50 before T50 T50 is reached closed stage is reached is reached Improvement of 14. Figure 10 Torque responses of the trajectories in Pareto optimal curve region 1. it was found that with the VGT 60% closed at the first stage. On the other hand.71 s 0.9%) (47.5 s (-6. Table 1 Comparison of the optimised strategies Requirements of transient response Fast torque Fast T50 and Highest response in High turbocharger turbocharger Fastest T50 the first 0.5 speed at 1. it is likely to cause a worse torque response at the later stage in the transient.5 second second second VGT position at 75% .5 second.5 second strongly affects the time required to reach higher torque (for example 90% of maximum torque rise). As the turbocharger speed at 1.8%) (0.1%) (35. because the achievable transient response was largely limited by the VGT setting at the first stage.5 s (10.4%) (-3.9%) 172 . The VGT position in the first 0.8%) 0. T50.5 krpm while the T50 can still be improved by 16. and the intervals between the optional VGT positions were 5-10%. the torque rise in the first 0. the steady state turbocharger speed was increased by 11 krpm (138. Thus. both the torque response and the turbocharger acceleration were worsened. and the turbocharger speed at 1. However. The turbocharger speed at 1. In order to avoid the constraints of using a feedback controller.4%) by fully closing the VGT. If the turbocharger speed at 1. The T50 was reduced by up to 47. fully closing the VGT during the transient resulted in high back pressure and low volumetric efficiency.The torque responses of optimised VGT actuator trajectories are compared with a single-step-change VGT actuator trajectory (fully closed before tip-in and 75% closed after tip-in). Compared to a single-step-change VGT actuator trajectory. a range of open loop VGT actuator trajectories were investigated to optimise the first 1.4%.5 s was not to be compromised. the T50 can be improved by 35.5 second. 4 CONCLUSION VGT turbochargers have been used predominantly on Diesel engines rather than gasoline engines due to the complexity of the mechanism and the lower exhaust gas temperature limit.5 second has been illustrated in this study.5s was also lowered by 6.1% although the torque rise in the first 0.5 seconds of transient operation during an engine load step change from 2 bar BMEP to full load. Therefore.9% and the turbocharger speed at 1.0L gasoline engine. the T50 improvement was limited at 33. larger improvement may be achievable if smaller time step and VGT position interval were used. A VGT turbocharger has been tested on a 2. Compared with fully opening the VGT at low load operation point before tip-in.5 second.5 s and the turbocharger speed at 1.5 second 173 . It should be noted that the duration of the three stages in the transient was 0. By pursuing the fastest torque build-up in the first 0.5%.5 s were worsened. shown in figure 11. The optimum calibration depends on the requirement of the transient operation. whilst the fuel consumption was maintained the same considering the measurement accuracy.5%.5 s can be improved by up to 0. Figure 11 Comparison of the optimised trajectories and the reference trajectory. A trade-off between the torque rise in the first 0.5 s. Wastegated Turbocharger and a Variable Geometry Turbocharger. The Turbocharger with Variable Turbine Geometry for Gasoline Engines. 68(2).6%) at the expense of lowering the turbocharger speed at 1. Jacob. If no turbocharger acceleration compromise was accepted. IC-Engine Downsizing and Pressure-Wave Supercharging for Fuel Economy.. L. Turbo-Cool Turbocharging System for Spark Ignition Engines. 1986. Part D: Journal of Automobile Engineering. TurboCentre. Ford Motor Company Ltd and Cummins Turbo Technologies Ltd.7 krpm (3. a model based transient control strategy may be required to improve the transient operation and to avoid overshoot. 2000. L. P. Donkin. Wenger. and Martin. Vol.5 second by 1... The Next Generation of Gasoline Turbo Technology. and the T50 can be improved by 0. SAE technical paper 2000- 01-1019 [2] Bauer. Shahed. M.. D...54 s (35. The turbocharger speed at 1. pp. A Comparison of Transient Vehicle performance Using a Fixed Geometry. M. Munkel. R.. U. 1992. 6 REFERENCES [1] Guzzella. S. G. pp. 114(3). Rodenhauser. C. The experimental work was carried out in the Powertrain & Vehicle Research Centre in the University of Bath.. H. pp. K.. and Gall.5%).25 s (16. SAE technical paper 2002-01-0161 [7] Gabriel. 96-103 [8] Capobianco. 1985. SAE technical paper 860104 [5] Singer. Compared with a conventional boost pressure feedback controller. a Turbocharger in a Small Displacement Gasoline Engine Application.9%). 2002. Advanced Variable Geometry Turbocharger for Diesel Engine Applications. Vol.5 second can be increased by up to 0.5 second by 3.. K. Journal of Engineering for Gas Turbines and Power- Transactions of the ASME. S.. P. Variable Geometry and Waste- Gated Automotive Turbochargers .Measurements and Comparison of Turbine Performance. while the T50 can still be improved by 0. Jaguar Land Rover Limited.-S.71 s (47. H. the T50 can be improved by 0. between the University of Bath. 1163-1175 [4] Lundstrom.-H. Vienna 2012 [3] Wang. S.8 krpm (6. 5 ACKNOWLEDGEMENT The data shown in this paper was from a collaborative research project.5 second was crucial. 553-560 174 . J. A. 2006. A. Groskreutz. Vol. Balis. and Yang...8%). R.5 krpm (0.4 Nm (10. U. 2007. SAE technical paper 850244 [6] Arnold. The VGT position in the first 0. and Schmalzl. The project was funded by the University of Bath’s EPSRC Knowledge Transfer Account and project partners.. 220(8). and Davies.can be improved by 14. Comparison of a Supercharger vs.4%). Proceedings of the Institution of Mechanical Engineers. M. H. G. and Gambarotta. and Slupski. because the achievable transient response was largely limited by the VGT setting at the first stage. M... MTZ. The authors would like to thank all the organisations in this project for the permission to publish this paper. 33rd International Vienna Motor Symposium 2012. which may fully close the VGT during transient. S.1%) at the expense of lowering the turbocharger speed at 1.5%). 1983.. Development of VFT (Variable Flow Turbocharger).. Hajek. Developments of Variable Area Radial Turbine for Small Turbochargers. 1999. and Buckland. Study of Variable Scroll Type Turbocharger (Determination of Shape of Scroll). North Branch: CarTech Inc.. and Martensson. Adachi.. 1989.. T. Nonlinear Exhaust Pressure Control of an SI Engine with VGT Using Partial Model Inversion. Y. and Roelke.. T. Okazaki. ed.. Kolmanovsky. I. K. 2010 [12] O'Connor. and Tomita. K. Turbo: Real World High-Performance Turbocharger Systems: S-A Design. and Kawakami. Atlanta. L. Boosted Gasoline Direct Injection Engines: Comparison of Throttle and VGT Controllers for Homogeneous Charge Operation.. E. SAE technical paper 891874 175 . SAE technical paper 880121 [13] Hirhikawa. SAE technical paper 831517 [16] Umezaki. and Busch. Kono. O.. J... M.. P. 1988. 2008. M. 49th IEEE Conference on Decision and Control. [11] Flardh. J. USA. 2002. Ogura. and Smith. Aerodynamic Effects of Movable Sidewall Nozzle Geometry and Rotor Exit Restriction on the Performance of a Radial Turbine. J. T. C. Variable Nozzle Turbochargers for Passenger Car Applications. A.. SAE technical paper 1999-01-1242 [15] Rogo. SAE technical paper 880120 [14] Kawaguchi. S.. G.[9] Lezhnev. 1988. R. SAE technical paper 2002-01-0709 [10] Miller. H E Ångström KTH . the exhaust flow is divided between two different exhaust manifolds with different valve timing. blowdown valve timing and scavenging valve timing.The exhaust energy utilization of a turbocompound engine combined with divided exhaust period H Aghaali. This architecture produces less negative pumping work for the same engine load point due to lower exhaust back pressure. However. The fuel-saving potential of this architecture have been theoretically investigated by examining different parameters such as turbine flow capacity. however. To reveal the full potential of this approach. increasing the effective flow area of the valves should be studied. CCGEx. In addition to the turbocompounding. To decouple the blowdown phase from the scavenging phase. Therefore. According to this study. there is a compromise between the turbine energy recovery and the pumping work. This is mainly due to the choked flow in the exhaust valves because this approach is using only one of the two exhaust valves at a time. To reduce the fuel consumption. There have been a few investigations into the DEP engine that all studies are about a turbocharged engine equipped with DEP [3-7]. The main drawback of turbocompound engines is the high _______________________________________ © The author(s) and/or their employer(s). This is called turbocompounding. A variable valve train system is assumed to enable optimization at different load points. Sweden ABSTRACT To decrease the influence of the increased exhaust pressure of a turbocompound engine. one of the main developments in internal combustion engines is waste heat recovery.2]. a new architecture is developed by combining the turbocompound engine with divided exhaust period (DEP). The aim of this study is to utilize the earlier stage (blowdown) of the exhaust stroke in the turbine(s) and let the later stage (scavenging) of the exhaust stroke bypass the turbine(s).2]. 1 INTRODUCTION The foreseeable shortage of fuel and the discussion of global warming have heightened the need for designing more fuel efficient engines. Internal Combustion Engines. This can be done by converting exhaust gases heat to mechanical work in a turbine. Many combinations of these parameters are considered in the optimization of the engine for different engine loads and speeds. the exhaust mass flow into the turbine(s) is decreased. this combination shows fuel-saving potential in low engine speeds and limitations at high engine speeds. DEP is another way to reduce fuel consumption that decreases mainly the pumping loss in engines [3-7]. as well. turbocompound turbines create pumping loss in the engines [1. Turbocompound engines have been widely investigated and it has been shown that they have fuel-saving potential in the range of 1-5% [1. Turbocharged engine with DEP has a limited fuel saving potential.Royal Institute of Technology. 2014 179 . Figure 1a shows that one port of each cylinder is connected to the blowdown manifold and the other port is connected to the scavenging manifold. However. the hypothesis will be analysed on one engine load point. The hypothesis is that a turbocompound engine combined with DEP is more fuel efficient because of the reduced pumping loss. The blowdown manifold feeds the turbine and the scavenging manifold bypasses the turbine. Figure 1b shows the timings of the exhaust and intake valves. This was done for the compressor shaft. This is a single-stage turbocompound engine that has been simulated previously with electric generator and motor instead of mechanical transmissions [1]. The turbocompound engine without DEP uses the ordinary intake and exhaust valves timings as the turbocharged engine. Kruiswyk [2] has concluded that elimination of EGR on a turbocompound engine improves the BSFC benefit by 1- 1. then more load points will be examined. GT-Power is an 1D fluid dynamic tool with engine flow models. the turbocharger turbine and compressor were disconnected from each other. the turbocompound engine was equipped with a VVA for each exhaust valve and two separate exhaust manifolds. the main advantage of DEP architecture is the reduced pumping loss. This engine incorporates a twin scroll turbine and no EGR system. 180 . Firstly. heat release rates based on measured cylinder pressures were used. lookup tables and empirical models for some components such as valves. This study was designed to evaluate the role of the DEP architecture on a turbocompound engine keeping the cylinder head geometry. and a combination of turbocompound and DEP. To model the combustion under different engine speeds and torques. This gave the possibility to keep the boost pressure constant for each load point while the turbine swallowing capacity was changed. which was chosen to keep the complexity of the initial study at a minimum. the turbocompound engine with DEP employs the blowdown and scavenging valves timings instead. In this paper. This paper gives preliminary results for fuel-saving potential of a turbocompound engine with DEP. the turbine efficiency and the boost pressure unchanged. Therefore.5% at high loads. variable valve actuation (VVA) and exhaust manifolds for DEP. On the other hand. Secondly. The turbocharged engine was then modified in the simulations to be equipped with a compound turbine. The studied engine is 11. the turbine energy recovery might be lower due to the reduced amount of exhaust flow into the turbine. It uses maps. in-line six cylinder heavy-duty Diesel engine. The size of the exhaust manifolds and their properties have been extracted from a previous study [10]. turbocompound engine. as illustrated in Figure 1. The engine specifications are provided in Table 1. extending previous research by combining a turbocompound engine with DEP is the topic for this analysis with the aim of reducing pumping loss while energy recovery in the turbine(s) is maintained. The baseline model for this engine was calibrated in a previous work [9] and it was used as the base for the simulations. However. 2 SIMULATION The simulations have been performed in GT-Power [8] on a heavy-duty Diesel engine for three different architectures. turbocharged engine.pumping loss. It has to be mentioned that the collision between the scavenging exhaust valve and piston is not considered in this study.7 liter. The turbine shaft was connected directly to the engine crankshaft by a continuously variable transmission (CVT). turbochargers and cylinders. Meanwhile. as well. Figure 2. The simulation approach was to minimize the break specific fuel consumption (BSFC) of the engine in different architectures. Engine label SCANIA DC1201 Emission class Euro 3 Max. 181 . Engine specifications. Table 1. a) The architecture of turbocompound engine combined with DEP. as illustrated in Figure 2. This has been performed for several engine loads and speeds. b) Variable valve timing including blowdown and scavenging exhaust valves.7 Bore [mm] 127 Stroke [mm] 154 Conn. Torque [Nm] 1900 Displacement [dm3] 11. rod length [mm] 255 Compression Ratio 18 IVO [°ATDC] 346 IVC [°ATDC] 142 EVO [°ATDC] 136 EVC [°ATDC] 359 Turbocharger Twin-entry EGR System No (a) (b) Figure 1. Chosen load points of the original turbocharged engine. Power [hp] 380 Max. any discrepancy from measurements is assumed to equally affect the models. In the simulation of the turbocompound engine. dissimilar pulsating flows will not disturb the results. the exhaust back-pressure was varied by scaling the swallowing capacity of the turbine. The goodness of the fits has been checked for all load points by calculating the coefficient of determination (R squared). Different exhaust back pressures will make altered turbine power and pumping work. the blowdown exhaust valve closing time and the scavenging exhaust valve opening time are varied. To find the optimum combination of these three parameters which gives the minimum BSFC of the last architecture. boost pressure and exhaust valves geometry. The combination of turbocompound and DEP has the lowest BSFC at this load point with the same turbine efficiency. For each load point. some parameters are kept unchanged and constant based on the original turbocharged engine for all three architectures such as: – Power output of the total system – Engine speed – Turbine efficiency – Turbine speed – Compressor efficiency – Compressor speed – The size of the intake valves – The size of the exhaust valves – Intake valve opening time – Intake valve closing time – Exhaust (blowdown in DEP architecture) valve opening time – Exhaust (scavenging in DEP architecture) valve closing time – Frictional mechanical efficiency of turbocharger = 100% – Mechanical efficiency of CVT = 100% Since the turbine efficiency is constant in each load point for the three architectures. At each load point.Since the current study aims at comparing the simulated turbocharged engine with the modified models. a model-based optimization has been performed using GT-POWER to minimize the BSFC response. The modelled parameters for each load point have been extracted and all load points are run again to be sure that the modelled parameters provide the same responses. The result shows that the turbocompound engine has lower BSFC than the turbocharged engine while the efficiency of the turbine is kept constant. This can be done by varying turbine mass multiplier in GT-Power. in addition to the turbine swallowing capacity. 182 . In the simulation of the turbocompound engine combined with DEP. 7 levels of the blowdown exhaust valve closing time and 7 levels of the scavenging exhaust valve opening time. They were all very close to 1 for all load points. rather than on the pulsating flow in the turbine. 245 cases have been considered that contain 5 levels of the turbine mass multiplier.1 Fuel-saving potential at high load and low speed Figure 3 shows the BSFC of three different architectures at a high engine load and low speed (A3 on Figure 2). The optimized exhaust back pressure of the turbocompound architecture which gives minimum BSFC has been calculated for each load point. The research has focused on the created exhaust back-pressure in different architectures. a full factorial design of experiment (DOE) has been run. Then. 3 RESULTS 3. however. To explain the potential of this architecture in fuel-saving. the trade-off between the surplus turbine power and pumping work shall be examined. The turbocharged 183 . this figure shows the relation between the surplus turbine power and PMEP. normalized PMEP by BMEP of load point A3 for three different architectures. In this load point the total engine power. turbocharged. consequently BMEP. Figure 4 shows the normalized surplus turbine power by the total engine power against the normalized pumping mean effective pressure (PMEP) by break mean effective pressure (BMEP) of the load point A3 for the three different architectures. positive and negative signs of the surplus turbine power indicate turbocompound and supercharged engines. Figure 3. turbocompound and turbocompound with DEP. it means that the engine is turbocharged. Figure 4. the total power of the system is the summation of net engine power and the surplus turbine power. respectively. On the single-stage turbocompound engine in which the turbine and the compressor are mechanically connected to the engine crankshaft. Fuel consumption of different engine architectures. Normalized surplus turbine power by engine power vs. is kept constant. When the surplus turbine power is zero. Therefore. The surplus turbine power is the difference between the turbine power and the compressor power. This point is marked along the line of the increased back pressure. Besides. In a turbocompound engine. For the turbocompound architecture at load point A3.9% “surplus turbine power/engine power” and almost -0. Therefore. This means -2. This figure provides a simple way to compare different architectures in terms of surplus turbine power and pumping work. the exhaust back pressure should be increased to create more turbine power.4% while “PMEP/BMEP” difference is about +1. Mach number through the exhaust valves vs. engine crank angle for three different architectures at load point A3. the pumping loss will be increased. Figure 5 shows the Mach number of exhaust flow through the valves for three different architectures at load point A3. In addition to the less exhaust mass flow. the optimum case of the combined turbocompound engine with the DEP architecture is marked around +4. however. it is preferred to move toward more surplus turbine power and higher positive PMEP.5% “PMEP/BMEP” is a result of the increased back pressure. the turbine energy recovery is less. The main concerning issue regarding the turbocompound engine with DEP architecture is the choked flow through the exhaust valves due to the small effective flow area of the exhaust valves. If we compare this point with the turbocompound one the difference between the “surplus turbine power/engine power” is almost -1. Figure 5.3% “surplus turbine power/engine power” is provided while -2. +6.6% which means turbocompound engine with DEP consumes less fuel. This is because the exhaust flow from the cylinders to the exhaust system is divided into two exhaust manifolds and the turbine is just fed by one of them. As a result. The increased exhaust back pressure has an optimum point where the BSFC of the engine is minimal. Although the turbocompound engine with DEP provides less negative PMEP compared to the turbocompound one.8% “PMEP/BMEP”.3%. in total turbocompound engine is clear benefit compared to the turbocharged engine. the optimum exhaust back pressure is lower. the exhaust mass flow through the turbine is less.8% difference in “PMEP/BMEP” compared to turbocharged one. On this figure. However.5% while “surplus turbine power/engine power” is changed +6. 184 . because the difference of “PMEP/BMEP” from turbocharged to turbocompound is -4. the turbine energy recovery is less in a turbocompound engine combined with DEP compared to the turbocompound engine.engine has zero surplus turbine power and almost +2% “PMEP/BMEP”. this falls suddenly to lower than the intake pressure and the exhaust pressure of the turbocharged engine during the rest of the exhaust stroke. As it is clear. the compression. reducing the time of choked flow in the turbocompound engine with the DEP architecture can reveal its fuel-saving potential more while longer choked flow can happen in some load points that will be discussed later.Apparently. This is shorter for turbocompound engine due to the higher exhaust back pressure. therefore. however. Figure 6 shows the cylinder pressure versus normalized cylinder volume by the maximum cylinder volume for load point A3 with logarithmic scales. the intake. The turbocompound engine with DEP has higher exhaust back pressure remarkably during the first period of the exhaust stroke (blowdown) than the turbocompound engine. In the turbocompound engine with the DEP architecture. the turbocharged engine is restricted by the choked flow during the beginning period of the exhaust valve opening. This leads to longer choked flow. logarithmic scales. The only difference appears during the exhaust stroke due to different exhaust back pressures. Therefore. Figure 6. choked flow happens at the scavenging valve as well. the pumping work is positive in the turbocharged engine while this is negative in the turbocompound engine. Due to small effective flow area at the beginning of the opening.volume diagram of load point A3 for three different architectures. the choked flow is shorter. To compare the cylinder pressures in these three architectures. 185 . the scavenging valve starts opening with an overlap. firstly the blowdown valve is opened with half area compared to the original engine. Thus the pressure ratio of the cylinder contents to the exhaust system is smaller. Cylinder pressure . the combustion and the expansion strokes are almost the same for all architectures because the boost pressure and compression ratio are kept constant. This means the effective flow area is much smaller while the cylinder contents have higher time-averaged temperature and higher time-average pressure for the blowdown period than for the scavenging period. These results differ for another load point and in some points this will be much worse in term of long choked flow. We should keep in mind that the provided results are at load point A3 where this combination is beneficial. Therefore. The turbocharged engine in this load point has lower exhaust back pressure than the turbocompound one. Before the blowdown valve approaches the closing. The cases with high engine speeds are declined by this combination. Figure 7. BSFC improvement of chosen engine load points for different architectures compared to the turbocharged engine.5% deterioration compared to turbocharged engine in A13. The optimum exhaust back pressure of the turbocompound engine is marked on the line. Constant parameters have been changed for each load point based on the original turbocharged engine. As shown in Figure 7. the combination of turbocompound and DEP does not improve BSFC in all load points and in some points it makes even very much worse condition like 4. Therefore. The combined turbocompound and DEP could not lie on the right side of the line where the trade- off between the surplus turbine power and PMEP is beneficial. However. The only points that gain from this combination are the cases at low engine speeds. This improvement is mainly at high loads. 3. there is no gain compared to the turbocharged engine and the turbocompound architecture is preferred in this load point.2% in term of BSFC. 186 . Figure 8 depicts normalized surplus turbine power against normalized PMEP of one load point which gains no improvement (A7) for three different architectures. As an example. The BSFC improvements in cases with medium engine speed are almost zero.This leads to almost zero pumping work while the blowdown pressure is utilized in the turbine. This is the competence of this combination. in spite of the optimization in the turbocompound and DEP architecture. the turbocompound improves the turbocharged engine from 0 to 2.2 Fuel-saving potential of all chosen load points Figure 7 shows the BSFC improvements of the turbocompound engine with and without the DEP architecture compared to turbocharged engine as a base for different load points in Figure 2. The line on the figure shows the direction of increased exhaust back pressure for the turbocompound engine. The turbocompound engine is beneficial on just high loads and at all engine speeds. Normalized surplus turbine power by engine power vs. normalized PMEP by BMEP of load point A7 for three different architectures. Figure 8. Figure 9. using a turbine with higher efficiency and modifying the exhaust valves would extend the working range of this combination. 187 . the combination of turbocompound and DEP gives improvements just in low engine speeds. However. This combination is not beneficial at the medium and high engine speeds. Figure 9 illustrates the summary of this study’s results on the engine map. The potential working range of the turbocompound engine and the combined turbocompound engine with DEP in term of fuel-saving. However. and Lindström. "Valve-Event Modulated Boost System: Fuel Consumption and Performance with Scavenge-Sourced EGR. the shown improvement is just due to the different exhaust back pressure and different exhaust flow through the turbine. 2012. and Angstrom. H. Divided Exhaust Period on Heavy-Duty Diesel Engines. 2012. A. S. "The role of turbocompound in the era of emissions reduction.3." SAE Technical Paper 2010-01-1222. doi:10. S.5 to 3% in low engine speeds if the cylinder head geometry. the combined turbocompound engine with DEP improves BSFC 0.4271/2010-01-1222. doi:10. D. "Divided Exhaust Period . H.. Royal Institute of Technology.. GT-SUITE. doi:10. and achieve better turbocharger matching or turbine selection. ACKNOWLEDGEMENTS The Swedish Energy Agency (Energimyndigheten) and Royal Institute of Technology (KTH) sponsored this work within the Competence Centre for Gas Exchange (CCGEx). J. several parameters and geometries are kept unchanged from the original turbocharged engine..and Fluid Dynamic Processes in Direct Injection Engines. Flow Theory Manual 7. Transient simulations of heavy-duty diesel engines with focus on the turbine. Johansson.. Sweden. http://www. J. Inc. N. S. F. turbine efficiency and boost pressure are kept unchanged compared to the original turbocharged engine. [9] Winkler. Divided Exhaust Period on Heavy-Duty Diesel Engines. and Angstrom. T.. Cronhjort.. 2013. P. B. This means that if we could change the geometries. get higher efficiencies of turbine and compressor. 188 . 2005.4271/2012-01-0705. 2013. The main limitation in this approach is the choked flow through the exhaust valves which could be an interesting topic for future work.. then the fuel saving potential should be higher and its working range would extend to higher engine speeds. Sweden. REFERENCES [1] Aghaali.4 CONCLUSIONS DEP architecture was introduced to a turbocompound engine to enhance the fuel- saving potential of the engine. 2010. [4] Roth. SAE 2013-01-2703. Grandin. "Divided Exhaust Period: Effects of Changing the Relation between Intake. [10] Gundmalm. this combination deteriorates the fuel consumption at medium and high engine speeds.. [5] Roth. "Valve-Event Modulated Boost System. J. and Angstrom....4271/2005-01-1150. Kruiswyk. This new approach is beneficial by separating the exhaust stroke into two periods and reducing the pumping loss. [7] Gundmalm. 2008. doi:10. M...com. Demonstration of Air-Fuel Ratio Role in One- Stage Turbocompound Diesel Engines. P. THIESEL 2012 Conference on Thermo. [3] Möller.. Keller.. H. Engines 6(2):739-750." SAE Technical Paper 2005-01-1150. Licentiate Thesis. and Sisson. The exhaust flow feeds the turbine during the blowdown period and it bypasses the turbine during the scavenging period. H." 10th International Conference on Turbochargers and Turbocharging.gtisoft. D. A. Licentiate Thesis. However.. and Becker. Stockholm. [8] G..A Gas Exchange System for Turbocharged SI Engines. 2012. [6] Gundmalm. C.. Royal Institute of Technology.4271/2013-01- 0578. Blow-Down and Scavenging Valve Area." SAE Int. [2] R. Cronhjort." SAE Int. As a preliminary result. Engines 5(2):538-546.. Stockholm. Since in this study. Further development of two-stage turbocharging systems for large engines E Codan. The expected engine results with the new turbocharging system are presented and discussed. Compliance with the emission limits is obviously a must for the commercial application of an engine. C Christen ABB Turbo Systems Ltd. t Temperature (K. The choice of basic design as well as the achieved performance are discussed. °C) C Compressor V298. In this paper some aspects of the design of ABB’s new generation two-stage turbocharging are presented. NOMENCLATURE be Specific fuel consumption (g/kWh) Subscripts D Diameter (m) ac Start of compression KJ Specific moment of inertia (kg/m3) rec charge air receiver n Engine speed (rpm) TI Turbine inlet p Pressure (Pa. ܸሶ Reduced volume flow (m3/s) HP High pressure Flow coefficient LP Low pressure Efficiency T Turbine VVT Variable valve timing C Mass ratio of trapped to stoichiometric air aC Specific compressor work: enthalpy head/peripheral speed squared Pressure ratio Specific mass (kg/m3) Time constant (s) 1 INTRODUCTION Main drivers in the development of large engines are emissions. operating costs and first cost. 2014 189 . bar) CI Compressor inlet pmax Firing. The NOx limits will be substantially reduced in the near __________________________ © ABB Turbo Systems Ltd. The focus is set on diesel engines with strong Miller timing and variable valve timing. Several studies have already been published showing the large improvement potential concerning engine efficiency. maximum pressure (bar) bmep Brake mean effective pressure (bar) Abbreviations T. showing the potential for steady state and transient operation. emissions and power density. Switzerland ABSTRACT Two-stage turbocharging is a logical development step for large combustion engines. but ease of service is very important for reliability and performance stability. The technical feasibility was proven. The first generation was developed for a target range of pressure ratio between 7 and 9 with equivalent turbocharger efficiency above 73% using whenever possible components taken from normal single-stage turbochargers. Operating costs are currently mainly defined by fuel costs. Forecasts were even exceeded. At the same time. Other possibilities are currently investigated on a research level but are not yet validated. Maintenance costs are a smaller component of the operating costs. On research engines. At this point in their evolution. has been successfully in use since 2010 (1) (2) (7). exhaust gas recirculation (EGR) or gaseous fuel. First cost of a power plant is influenced by many different factors. Engine efficiency gains and emissions reduction clearly beyond the potential of any single-stage system have been demonstrated. This temperature reduction has a very positive effect on the engine’s efficiency and its raw NOx-emissions. Engine efficiency will become ever more important. Two system frame sizes have been released for serial production. The SOx emission will also be limited.future leading to the necessity to adopt after treatment (SCR – Selective Catalytic Reduction). the high turbocharging efficiency achievable with two-stage turbocharging allows the efficient delivery of the required air pressure and additional power due to the improvement in gas exchange work.1 First generation Power2® The first generation of ABB’s two-stage turbocharging solution. about 50 Power2 systems have been taken into operation. 190 . Since these components have not been originally designed for the requirements of a two-stage system. 1. The Miller process is here a very important because it allows the knock limit to be shifted. The situation on gas engines is different. it is not expected that power density will be considerably increased. which helps to keep a high efficiency over the whole life cycle of the power plant. which are monitored with the introduction of control parameters such as the Energy Efficiency Design Index (EEDI). Until today. These studies show that there is substantial improvement potential for engine performance due to the reduction of the process temperatures offered by the Miller process. values of mean effective pressures of up to 40 bar have been explored. High-pressure turbocharging in combination with the Miller process can offer significant improvements for all the aspects mentioned above. engine efficiency correlates with CO2 emissions. On diesel engines. This potential has been shown in several studies performed by means of advanced simulation tools and confirmed by engine tests (1) (2) (3) (4) (5) (6) (7) (8). and the power gap with diesel engines to be closed. but the development effort and the challenges of keeping efficiency and emissions on good levels limit the commercial feasibility. given by the product of mean piston speed and mean effective pressure. The ABB A100 turbocharger family has proven its potential for fulfilling today’s emission limits efficiently without additional measures. it is clear that some further potential would be available with a specific new development. front runner systems could be operated for more than 10. Additionally. but for sure one of them is the power density. Even more potential can be disclosed by applying two-stage turbocharging and variable valve timing.000 hours. because the knock limit is the main limitation for the power density of highly turbocharged lean-burn gas engines. leading to the necessity to use low sulfur fuels or devices able to reduce the SOx content of the exhaust gas. Power2. which have increased at least by a factor of five in the recent years. But for the global optimization many other aspects must be considered. For the new generation Power2 a wide range of pressure ratios from 8 to 12 was set as design target. without touching the hot parts. and a cartridge group. service friendliness was considered from the very beginning. Another contribution to engine downtime reduction is the inspection of turbochargers by endoscopy. The designated goal was to reduce service time of the complete two-stage system below the reference value of current single-stage turbochargers. equivalent turbocharger efficiency over 75% with an intercooler temperature of 70°C was set as another requirement.2 Optimum performance The overall performance of the turbocharging system highly depends on the design of its core components. In this range. After removal of the air casings it is possible to perform a quick exchange of the cartridge from the cold side of the turbocharger. 2. Consequently. consisting of air and gas casings. the target pressure ratio might be about 8. The idea behind this concept is that the turbocharger has an outer shell. If we take into account the trade-off between efficiency and emissions. a single figure compression ratio would not be the right solution. compressor and turbine stages.1 Ease of service Engine availability is a key factor for achieving optimal economic performance. were defined for the design. which contains the entire interior of the turbocharger. During development of the new turbocharger generation.2 POWER2 – SECOND GENERATION Looking at the design of the two-stage turbocharging system for four-stroke engines under the aspect of fuel efficiency only. only the air inlet casing together with the insert wall of the compressor needs to be removed. For the second generation Power2 turbochargers the extractable cartridge concept was developed (Figure 1). 2. Keeping in mind the overall performance to be achieved. All other interfaces with the engine as well as the insulation remain untouched. The cartridge consists of rotor. In order to exchange the cartridge during service. 191 . Figure 1: Extractable cartridge concept Today’s ABB turbochargers for medium-speed engines feature the cartridge group. which enables quick exchange of the cartridge. A dedicated tool was developed together with the extractable cartridge concept. which allows a status check of the components without removing any part of the turbocharger. which will be illustrated below. the time required for service work needs to be reduced to a minimum. bearings and directly related casings. as well as the requirement to improve power density and altitude capability. additional objectives. which are dimensioned for the specific thrust on different applications. The location and shape of the injection holes for water cleaning have been optimized by means of two-phase flow CFD simulations. Retaining this concept in the two-stage system was a key priority. 2. In order to optimize the bearing losses different thrust bearings have been developed for the HP turbocharger. Therefore. which leads to a 75% reduction of the air leakage compared to state-of-the-art labyrinth seals. which is considerably higher than in a conventional turbocharger.All connections in the casings of a two-stage system have to guide a volumetric flow. as the a) initial design. it is obvious that the dimensions must be kept as small as possible. but also Figure 2: LP gas inlet casing reduced dimensions. especially for engines burning heavy fuel oil is another issue. Performance stability in operation. b) optimized design example in Figure 2 shows. This is a highly undesired effect of the turbocharging system. The high-pressure level of the HP turbocharger as well as the presence of two turbocharger shafts would lead to a considerable increase of the air leakage into the engine oil system (blow-by). all the casings have been optimized for low pressure losses without exerting negative influence on the stage operation by means of CFD calculations. For this reason very high specific flow was set as target for all components. Since one turbocharger is replaced with two. a) b) The result is not only a better system efficiency. Regular turbine and compressor cleaning are effective measures to assure this performance stability. so a new sealing concept was developed.3 Compactness Almost every engine has its turbocharger mounted on it to form a complete unit. In order to illustrate the achievements the operating envelopes of the new developed components was plotted in the diagrams showing the historical development of ABB compressors (Figure 3). 6 2009 1-stage compressors C 2009 2003 5 2004 2003 1996 LP-compressor 1996 2012 4 1992 1978 1989 1983 1970 3 1954-1964 1946 2010-2013 2 HP-compressor 1924 1 V298 [m3/s] Figure 3: Compressor development 192 . The compressor map is plotted in Figure 4 in comparison with the compressor stage of the A100.The position of the HP compressor. The reasons are the superior performance of the axial turbines at high specific flow and moderate expansion ratio as well as the better suitability for HFO application. The LP stage maps of A100-M (single-stage) operates at high expansion ratio and and Power2 800-M LP (two-stage) a high efficiency is required. On the turbine side the situation is similar. An important decision was to adopt axial turbines for both stages to cover the application on medium-speed engine with power between 3 and 10 MW per turbocharging group. The first prototype of second-generation Power2 can be seen in Figure 5. Thanks at equal impeller wheel diameter to the lower exhaust gas temperature after the first expansion. which reaches much higher pressure ratios but at lower specific flow. Figure 5: Power2 850-M 193 . because the development was necessary already for the first generation. Very high specific flow at moderate expansion ratio is required Figure 4: Compressor performance for the HP stage. Completely new is the LP compressor. A second HP compressor stage was developed for applications requiring high overall pressure ratios. which reaches a range of pressure ratio closer to that of single-stage machines but with a very high specific flow. working at high specific flow and moderate pressure ratio has already been shown (5). the specific flow requirements are lower. Potential for power increase had not yet been considered at this stage. The green area represents the range of results expected for higher levels of bmep and turbocharging efficiency. 110 110 be NOx 1-stage_be_rel 2-stage_be_rel [% ] [% ] 105 100 1-stage_NOx_rel 2-stage_NOx_rel A 100 90 B 95 80 C D 90 70 85 60 80 50 4 5 6 7 8 9 10 11 12 C* Figure 6: NOx. It can be seen that the Volume 50 pressure increase with single- stage turbochargers. It was assumed that the volume is proportional to 100 Volume the third power of the linear Volume Volume dimension.2) and exhaust gas temperature (tTI = 520°C). The turbocharger have only been scaled in order to work in every case with a comparable loading. An example is given in Figure 6 (5) where the effects of the increase of boost pressure are shown for applications with single-stage and two-stage turbocharging for a medium-speed diesel engine. [%] HP Compressor diameter represented by the compressor 150 Volume wheel diameter. that for the design point there is a potential for improving the specific fuel consumption and that the NOx emissions are almost linearly decreasing with the boost pressure. The change to two- stage turbocharging at the same Figure 7: Turbocharger dimensions 194 . It can be seen.3 POWER2 APPLICATION Possible thermodynamic achievements with high-pressure turbocharging have already been shown in several theoretical studies. Every point of the curves represent an engine setting for constant output (bmep = 25 bar). The diagram 200 in Figure 7 shows the evolution of LP Compressor diameter the turbocharger dimensions. In order to better illustrate a possible development path. the gas exchange cycle is improved due to the increased turbocharging efficiency. and the points C and D for two-stage turbocharging have been chosen and more detailed simulations have been performed with performance maps of real components. Furthermore.and specific fuel consumption (be) over pressure ratio (C*) for single-stage and two-stage turbocharging systems The diagram was built with simulations with partly idealized components and considering the design points only. the points A and B for single-stage turbocharging. case A to case B requires an increase of the 0 Case A Case B Case C Case D dimensions. These effects are due to the positive effects of reducing the cycle temperature by means of the Miller process on the thermal efficiency as well as on the NOx formation. firing pressure (200 bar). air-fuel ratio (C = 2. mass the combinations of efficiency and turbine area flow rate for a free running have different effects at part load: the higher turbocharger 195 . The operation without variable valve timing is considered unfeasible. For the FPP curve the degree of speed 5 variation is increased with the full load bmep. but further increasing the pressure ratio.1 Part load behavior A turbocharging system without control is matched and operates well at constant engine speed.1 Case study medium-speed engine The engine model is the same that was used for the simulations in Figure 6. some boost pressure is TC = 0. Configurations C and D are realized with two-stage turbocharging. but by a consistent design taking into account the air density it can be further assumed that the volume of the turbocharging system is proportional to the volume of the turbochargers. 3. the valve timing is chosen for optimum filling. For the configuration A the turbocharger is a TPL-C.e. as realized with the ABB product VCM® (Valve Control Management).65 Series7 aspirated engine. i. 3.e. They are calculated for a free running turbocharger with constant 1 temperatures and efficiency. This configuration was abandoned with the introduction of the IMO regulations on NOx emissions (IMO Tier 1 and Tier 2). The pressure ratio of Point C is below the design range of Power2 and is considered only for comparison. The situation today with single- 3 stage turbocharging is not very much different. The configuration B goes to the limits of single-stage turbocharging. brings a reduction.e. i. because it is 6 just above the range of operation as natural TC = 0. but Figure 8: Pressure vs. Several studies have been performed in the past to evaluate the possibility to run on FPP (10). leading to comparable dimensions as case B. For marine applications with highly turbocharged engines the fixed pitch propeller curve (FPP) is a challenging requirement.pressure ratio implies a further considerable growth of the system dimensions.75 Series8 required. In Figure 8 it is shown how the pressure curves 2 over mass flow rate changes with efficiency and pressure level. a medium sized medium-speed diesel engine with an output of 5 MW per turbocharging group. In (10) a requirement parameter was defined in order to quantify the degree of speed reduction for a given operating line: nbmep10 (1) n10 1 n100% Charge air pressure [bar] A bmep level of 10 bar was chosen. without any Miller effect. A diagram based on experience first published in (12) shows that the level of full load bmep with 4 allows to run the FPP curve without control is about 23 bar. i. utilizing a turbocharger A100-M with radial turbine. as from case C to case D. A two-stage system has additional piping and a second air cooler.1. because the engine speed reduction at constant torque is always accompanied by a boost pressure reduction. but included in some cases as demonstration. For every pressure 0 20 40 60 80 100 Mass flow rate [%] level and efficiency the turbine area can be chosen for delivering the target pressure. The configuration of Point D includes the combination with a continuously variable valve timing. 3.the pressure and the efficiency. beCCCC D – C 7.5 T . in all cases the minimum air to fuel ratio is below the acceptable limits for operation with HFO (C = 4.3 (10).5 0. 1.100% K teil C (2) TC .5 350 The possibilities to improve the situation in case A are 1 140 300 200 well known: some NOx-Emission overboosting.9 range for running FPP without control.4 Kteil Necessary range 1.9).6 0.100% T .5 2-st. but Figure 10: Propeller curves without control 196 .3 1-st.5 1-st.1. Case B A beAAAA B beBBBB C – C 5. This 1. A therefore the part load parameter is 1. 1.5 650 1.8 TC.1 almost only a function of the full load efficiency (Figure 9).2 Engine performance on the propeller curve The main simulation results can be seen in Figure 10. for full load bmep = 25 bar about 1. The position of B 1. 0. 2 Air/fuel equivalence ratio 400 C 1. efficiency is higher.3 the part load efficiency by the ratio of the turbine flow coefficients. pulse [%] be 120 100 [%] turbocharging or the 110 50 introduction of control Specific fuel consumption options like exhaust gas 100 0 waste gate and air bypass.7-1. beDDDD gives the best part load results without control.5 parameter Kteil should be well above = one.5 1. the steeper is the decrease of pressure at part load.2 parameter cannot change very much. As predicted by the diagram in Figure 9. 90 -50 In any case a compromise is 80 -100 needed for the full load point 0 20 40 60 80 100 120 leading to a higher specific Engine load [%] – C 4. An equivalent part load efficiency can be defined multiplying 1.5 550 [°C] Figure 6. It is evident that operation with pressure above 6 bar and high efficiency is very challenging.0 the points for the four configurations C D considered are all below the required 0.0 2-st.100% even though with two-stage turbocharging the equivalent part load Figure 9: Part load parameter vs. A control is turbocharger efficiency required in all cases. – C 5. The specific fuel Turbine inlet gas temperature 4 600 consumption is consistent t TI with the reference values in 3. the differences are 3 500 maintained in the range 50 2. The different effects of efficiency and turbine area on the part load behavior of a turbocharging system have been taken into account in a part load parameter (10): TC .7 0.25 T For a good part load behavior the 1. fuel consumption. a part load 130 150 NOx optimized specification.5 450 to 100% engine load. But in the low load range. complexity A – C 4.5 2-st. More care is required. which is optimized for filling.0 2-st.9. In the load range 0 20 40 60 80 100 Engine load [%] between 0 and 30% the cylinder pressure is below 1 bar. because due to the high-pressure ratio the compressor map width is reduced and the negative effect of a reduced turbocharging efficiency at full load is larger. NOx-Emission The simulation results are 130 150 NOx excellent over the whole be 120 100 [%] operating range (Figure 11). This can be well compensated at high engine load by the turbocharging system. 6 Case D prec showing the charge air pressure prec and Case D pac 5 the cylinder pressure at compression start pac for the cases A and D on a 4 constant engine speed line. Extreme Miller timing means a reduction of the filling efficiency down to values of about 50%. VVT and reliability issues with beCCCC beDDDD HFO operation.5 350 possible to control the air. 80 -100 Variable devices on the gas 0 20 40 60 80 100 120 side would represent Engine load [%] additional costs. Since some Miller is present. which can be matched for the necessary increase of the charge air pressure.5 1-st. from idling up to about 30% load this is not possible. the charge air pressure is 2 considerably increased. the two 3 pressures are equal. [%] 110 50 Another big advantage of this Specific fuel consumption solution is that the 100 0 turbocharging system itself 90 -50 does not need any control.5 550 [°C] timing. D – C 7.1. valve timing variability can improve the situation. Completely different is the 4. [bar] Case A pac This issue is illustrated in Figure 12. Even in case C the valve timing variability is not sufficient to reach a minimum value of C = 1. 1 because the lower temperature allows to reduce the pressure in the design point 0 for the same mass. 1 300 140 200 fuel ratio in a wide range. beBBBB C – C 5.3 1-st. which gives a Figure 12: Effect of reduced comparable situation as a conventional cylinder filling 197 . other control devices in the two-stage turbocharging system are not considered an attractive solution.5 Turbine inlet gas temperature 650 situation in case D with the 4 600 tTI adoption of variable valve 3.5 450 higher than in the other Air/fuel equivalence ratio cases and changing the 2 400 C cylinder filling makes it 1. The available charge 3 500 air pressure is considerably 2. but the cylinder pressure is even lower than in case A. because the available exhaust energy is p 7 Case A prec too low for an effective turbocharging. but may not be sufficient. beAAAA B – C 5. Case D with variable valve timing 3. In the case D with strong Miller.3 The Miller effect at very low load Further considerations lead to the conclusion that variable valve timing for case D is very effective and necessary. All this can be Figure 11: Propeller curves – avoided for case D.the situation needs to be improved with the same measures as in case A. In the case A. 6 can be accepted. The latter can be expressed with compressor characteristics.2 at full load is shown for the 1 cubic propeller curve (n3). the resulting rule is: The higher the nominal charge air pressure. which controls the time required by the turbocharger to reach a new steady state condition. turbocharger and driven system.engine running at 5000 m altitude. engine speed – bmep = 10 bar In case A the possible speed reduction at C ≥ 1. In the cases with Miller timing the 2. In the cases B and C the cubic propeller curve is approached but not yet reached. on many other parameters like load profile.2 for keeping the air-fuel Case B equivalence ratio C at a 2 Case C constant value of 1. leading to the formula: 4 = ∙ ∙ (3) ∙ 198 . mechanical inertia of engine. In the diagram 1. thermal capacity of all system elements. Achieving a comparable air to fuel ratio at idling as on a conventional engine would require a pressure ratio about 2.1. two-stage turbocharging with variable valve timing. the 0 20 40 60 80 100 Engine speed [%] double cubic propeller curve (2∙n3) and the quadratic Figure 13: Air-fuel equivalence ratio propeller curve (n2). engine governor settings and other controls. the larger the speed range that can be covered. determining the available steady state air excess and consequently the possible sudden load increase and the time constant of the turbocharger. as will be shown in the following.4 the required speed reduction starting from bmep = 25 bar 1. 3. In case D the propeller curve can be run with a margin for 50% overload and even the double propeller curve could be run. With the concept of case D.6 is reached.4 inlet valve timing is controlled C n2 2∙n3 n3 Case A 2. besides on the thermodynamic characteristics of engine and turbocharging system. inertia of the gas in the turbocharging system. which is probably only feasible with mechanical supercharging.4 Operating range with variable valve timing High-pressure turbocharging with variable valve timing offers even the possibility to enlarge the operating envelope beyond the propeller curve. the dominating influence for a fast load response is.5 Transient operation The transient operation of a diesel engine is extremely complex. But. its contribution to the load acceptance depends on two aspect: The part load performance. vs. if a value of C = 1. if the compression ratio is not high enough. The lines in Figure 13 represent the possible speed reduction at bmep 10 bar in the different cases. In this condition it may be impossible to start the engine. 3. because it depends. Looking at the turbocharger only. It makes much more sense to change the valve timing. how fast and how much the exhaust gas expansion energy in the turbine can be changed.9 until the 1.1. The time constant of a turbocharger is defined as twice the rotational kinetic energy of the rotor divided by the shaft power.9 is very limited.8 Case D condition of maximum filling 1. 5 Figure 14. The results can be seen in constant C = 1. because the Miller effect reduces the power margin at the beginning of the ramp and the turbocharger has a larger time constant. depending on the volume between the stages. In addition to this the HP turbine accelerates faster. defined as the moment of inertia divided by the fifth power of compressor diameter and of the compressor diameter.The time constant is a linear function of the inertia coefficient KJ. In case D with constant valve timing the initial power increase is further lowered. In the attempt to 80 B find a representative case.49 1.5 made usually with a linear ramp. 199 . Table 1. With comparable Miller effect and two-stage turbocharging (Case C) the load acceptance time is reduced by about 20%. On 40 Case A Power PIV 4. The faster acceleration of two-stage systems can only partially be explained with the lower time constant. simulations have been performed 60 for a load increase from idling to 100% at constant speed. Since the compression temperature is very low. In the 0 simulations an ideal controller -5 0 5 10 15 20 25 30 Time [s] was used. It must be remarked here that case D with constant valve timing represents the behavior of the system under the assumption that the combustion works well. This problem disappears by applying variable valve timing. a sudden increase of the expansion ratio for the HP turbine results. 20 Case C 2-st Power PIV 5. but this is overcompensated by the superior transient performance of the two-stage system with lower inertia. which can change the [%] 100 influence of different system D A C parameters. it is possible that the long ignition delay considerably reduces the combustion efficiency. therefore for comparison it can be defined for a constant pressure ratio or a constant engine load.34 Transient loading of large Engine 120 engines can follow different output D+VVT patterns. the remaining enthalpy after expansion in the HP turbine.0 power plant applications this is Case B 1-st Power PIV 5.82 1. because it can use initially a very large part of the exhaust enthalpy.0 Case DVVT PIV 7. whereby the LP turbine receives with some delay.64 1. When the gas pressure in the exhaust manifold is increased. The time constant depends on the operating point. Case B requires about 22 s.12 LP stage 1. which is able to keep the air fuel ratio on a constant Figure 14: Load ramp 0 – 100 % with value. the specific work and the specific flow of the compressor stage. In Table 1 the time constants are given for the four cases at 70% engine load. In all these cases with fixed valve timing the time to full torque is considerably higher than in the reference case.5 whose gradient is dictated by a Case DPower PIV 7. Time constants of the considered configurations Time constant Case A Case B Case C Case D [s] HP stage 1.57 2.0 + VVT Power specific critical load range.70 1. but applying variable valve timing this time is reduced to less than 6s. Under the defined boundary conditions case A reaches 100% power in less than 10 s. On the denominator side of the fractions are the density at compressor inlet.48 TC system 1. 4 3%. The four cases A.6 For the case D’ the dimensions of the stages are changed: The Series1 CLP/CHP = 2 linear dimensions of the HP Series2 CLP/CHP = 1. those of the LP turbocharger Range for optimum efficiency must be increased by 9%.2 design point is improved by about 2 points. 2 0.87 =2 [bar] against case D’. comparing the results of case D (C. optimum values between 1.1. C.87 = 1. which is only marginally lower in the design point.3. But since the pressure Pressure ratio split ratio of the LP stage decreases much faster with engine load 0. case D 0 20 40 60 80 100 Engine load [%] has a better efficiency in the middle load range (Figure 15).3. it runs faster speed -2 [-] and accelerates faster.LP/C. The turbocharging efficiency in the 1 0. 3.3 MW per turbocharging unit was investigated by means of simulation. CLP/CHPnTLHD 510.5 0.5 4 Interesting is also the comparison Charge air pressure of the transient operation (Figure 2 16). B.8 Turbocharging efficiency 2) with a case D’ with C.2 and 1. But the 0.3 6 1. It has already be shown (3)(5) that a value of 2 gives an efficiency.2 Case study high-speed engine In this case a high-speed diesel engine with a power output of about 1. By reducing the LP stage 1 0 pressure ratio the HP stage Normalised HP TC speed makes more work. This implies some over-boosting in the design point with wide compressor maps and allowance for higher exhaust gas temperatures. 200 .3 turbocharger can be reduced by 1. Typical for high speed engines is strongly reduced overlap and a wider operating envelope to be run without control.5 0.5 -4 consequence is that the delay of LP TC speed the LP stage is increased and the -6 pressure ratio for full power is 0 -8 reached later. With typical values of these temperatures. The thermodynamic effects of the change have been studied. D are defined in the same way as for the previous case study. but leads to a smaller LP turbocharger.5 results for the ratio C. The best system -5 0 5 10 15 20 25 performance is achieved when Time [s] both stages can contribute and Figure 16: Effect of pressure ratio this is realized better in the split on transient operation reference case. In Figure 15: Effect of pressure ratio the part load region case D gives split on steady state operation an improvement of 3% for the air-fuel ratio and 13 °C exhaust 2 8 p gas temperature reduction CLP/CHPnTLHD 510.HP = 2. All turbochargers involved are with radial turbine.6 Effect of the pressure repartition The repartition of the pressure ratio between the stages which gives the best efficiency in the design point depends on the ratio between intercooler and ambient temperature.LP/C.5 0 than that of the HP stage.LP/C.HP = T 1.HP. 193 1.2 Transient operation The time constant of the resulting turbocharging systems for the high-speed engine cases are shown in Table 2.3. B.3.min = 1.tot = 7.3 201 .98 1. thus.37 Transient simulations have been carried out in the same manner as for the medium-speed engine (Figure 18).min = 1.7. However. Table 2. Cases B and C with roughly the same Miller effect 60 Case A show the same level of initial power 40 Case B step. While C 94 in case A the medium-speed engine D is being operated at a pressure ratio of 4. Time constants of the considered configurations at 70% engine load Time constant Case A Case B Case C Case D [s] HP stage 1. Figure 17: be-NOx trade-off diagram At the nominal load point a very similar pattern of specific fuel consumption reduction as in Figure 6 can be observed. 3.1 The engine performance at constant speed 100 A Figure 17 shows the full engine be power performance of cases A. Case D VVT Figure 18: Load ramp 0 – 100 % ramps up in less than 4 seconds. C [%] B 98 and D in a trade-off diagram of fuel consumption and NOx emission.2. the Case C subsequent ramp up is dominated by 20 Case D the TC system time constants. NOx emission is reduced by 40% compared to the nominal point of case A.83 TC system 1. The nominal operating points feature 96 constant levels of air/fuel ratio and maximum cylinder pressure.37 1. 90 the distance between the cases A 0 20 40 60 80 100 NOx [%] and B is reduced in the high-speed engine case. the HS engine at bmep 25 92 HS Engine bar without Miller requires a MS Engine pressure ratio of about 4. At a total pressure ratio of c.12 0. The tendency of NOx emission reduction with increasing Miller timing turns out to be more pronounced in this case compared to the medium-speed example. with constant C. After this initial step. The initial increase of output [%] engine power is reduced from case A 100 to case D according to the reduction of the charging efficiency (Miller 80 effect).2.53 1.69 LP stage 2. the controller limit was extended corresponding to a lower limit of excess air/fuel ratio Engine 120 of C.3. Case Case D + VVT D with VVT clearly outperforms the 0 reference Case A: while the reference -5 0 5 10 15 20 25 30 case reaches full engine power within Time [s] roughly 7 seconds. since the use of heavy fuel oil can be excluded for this type of engine. Therefore. because the added value is too small. +++ Transient performance 0 .0 acceptance in two steps. + ++ increase Case B (A100) represents the last step in the development of single-stage turbocharging. In this operation mode the advantage of 0. the HP stage.min = 1. Evaluation case studies Case A Case B Case C Case D Compactness 0 . The last generation of ABB A100 turbochargers was developed to exploit the full potential of the single-stage turbocharging technology. In this paper different aspects of the system performance have been studied for two specific engine type cases. Table 3. Case D. If the load step is Case D + VVT smaller. 202 .Figure 19 shows the transient Engine 1.overall Figure 20: Weighted value function for The results can be summarized as single. is not considered attractive. +++ NOx emissions 0 + + ++ Potential for power 0 . -. Case C. replacement of single with two-stage turbocharging at much higher pressure ratio in combination with Miller and variable valve timing gives access to a large improvement potential. The requirements set by case B in the study can be fulfilled with excellent performance.6 more pronounced than in the preceding case. Figure 19: Engine load acceptance in island mode with constant c. - Engine efficiency 0 + ++ +++ Part load 0 + . The reason is that Case A 0. where the value of single and two-stage turbocharging systems is not optimal.and two-stage turbocharging in Table 3. . which reacts 0.4 in the long ramp both Case B turbochargers must contribute to 0.3 CONCLUSIONS Value 1-stage A value diagram was presented in 2-stage (3) showing qualitatively that there is a range of pressure ratio. C. can provide the 0 5 10 15 20 25 30 Time [s] necessary pressure change also without contribution of the LP stage.0 much faster. It gives an improvement in efficiency and emissions (IMOII) using the potential of the single-stage technology. replacement of single with two-stage at constant pressure ratio.2 speed behavior of the engine operated in [-] island mode during a load 1.2 Case C + VVT the load step.8 two-stage turbocharging is much 0. Aufladetechnische Konferenz. U. E. Emissions – a new challenge for turbocharging. 14. (12) Codan. B. 26th CIMAC World Congress in Bergen (N). Within the design range of second-generation Power2. Potential des Caterpillar MaK 6 M32 C mit zweistufiger Abgasturboaufladung.ABB’s second-generation Power2 and VCM are innovative products designed for the efficient exploitation of the two-stage turbocharging technology on large engines. (4) Codan. 2011. 15. 2010. 2007. Combining dual stage turbocharging with extreme Miller timings to achieve NOx emissions reductions in marine diesel engines.. Trapp. Dresden (D). 15th CIMAC World Congress. M. 1983. D. 26th CIMAC World Congress in Bergen (N). Hallbäck & A. Zweistufige Hochdruck- Turboaufladung für Gasmotoren mit hohem Wirkungsgrad. 25th CIMAC World Congress in Vienna.. 5. Applications and Potentials of two- stage Turbocharging. (www. Rickert. 203 . (7) Ruschmeyer. 2009. Part-load operation of very highly turbocharged four-stroke diesel engines. Ch. A.. G. J.com). 2010. Klausner.. (5) Codan. An additional asset of two-stage turbocharging systems is the possibility to compensate altitude without any need for de-rating.. K. (3) Codan. 6. E. M. the high pressure ratio and the high efficiency are important: The engine designer is free to set a bmep target in the development of new engines without considering any limitation from the turbocharging system.cimac. S. Müller. (10) Codan. & Rettig. Design and first application of a two- stage turbocharging system for a medium-speed diesel engine. In the study the engine output was kept constant for a better comparability. (8) Mathey. J. Dresden (D).... E. E. & Schlemmer-Kelling. (11) Meier. Ch. F. 27.. Optimierung des Aufladesystems und Betriebsverhaltens von Grossmotoren durch Computersimulation. Ch. Anforderungen an Aufladesysteme für zukünftige Grossmotoren. Aufladetechnische Konferenz. 1997. 2010. Recommendation Nr. the part load and the transient behavior were comparable or better than for case D. Aufladetechnische Konferenz. Dresden (D). Augsburg (D). C. 2010. Ch. Lang & Ch. Mathey. T. Frankfurt am Mein (D). 2007.. & Vögeli..5 was considered. 26th CIMAC World Congress in Bergen (N). Gianoglio. 2010. In a further extension of the present work a case E with pressure ratio 8. 26th CIMAC World Congress in Bergen (N). 1993. Variable Valve Timing – A necessity for future large diesel and gas engines. Turbocharging Efficiencies – Definitions and guidelines for measurement and calculation. Conseil International des Machines à Combustion. Aufladetechnische Konferenz. Paris. two-stage Turbocharging – Flexibility for Engine Optimization. REFERENCES (1) Raikio. (9) CIMAC. E.. Dresden (D). (2) Haidn. 16. & Delneri. (6) Millo.. In all cases the fuel consumption. & Mathey. applying either more Miller at constant power for emission reduction or a power density increase to bmep = 30 bar at constant Miller effect. Hjort. Aufladetechnische Konferenz. Mathey. Austria.. E.. Worldwide legislative requirements and customer desires for improved efficiency whilst maintaining vehicle drivability and performance are challenging automotive manufacturers to investigate innovative concepts for air charge delivery. This paper provides an outline of an innovative charging system concept developed to support future Diesel engine applications. Details of the advanced control mechanism and interaction with the electrical system including load management will be reviewed.Electrical supercharging for future diesel powertrain applications P Newman. N Luard. UK C Rochette. R Jackson Lotus Cars Limited. Under the UK Government funded Technology Strategy Board project ‘Provoque’ the partners have collaborated to develop a compound charging system incorporating a standard LP VGT turbo and 48V powered electric supercharger. Jaguar Land Rover Limited. S Jarvis. S Richardson Powertrain Research & Technology. 2014 207 . ABBREVIATIONS BSG Belt Starter Generator EGR Exhaust gas recirculation JLR Jaguar Land Rover NEDC New European Drive Cycle 1 INTRODUCTION 1. the drivability and performance capability. Application of the innovative charging concept to a premium vehicle will be discussed with indication of both. M Criddle Valeo Air Management UK Limited. Lotus and Valeo. UK T Smith.1 Global Trends and Diesel Engine Boosting System Increased customer awareness of environmental issues in the worldwide automotive market has driven Governments and automotive manufacturers to focus on developing technology to support vehicle CO2 improvements as well as maintaining the historically challenging legislated criteria emissions [1]. D Lee. ____________________________________ © Jaguar Land Rover. UK ABSTRACT The desire to minimise fuel consumption and corresponding vehicle CO2 emissions for future powertrain applications drives the need for advanced charging systems. as well as the ability to support an aggressive down-speeding concept facilitated by the electric supercharger to further improve efficiency. Validated analytical results for the hybrid compound charging system will be presented to show the application over both current legislative and real world replicating drive cycles. VW [11] BMW [12] can be all seen in literature. it is projected that a significant reduction in vehicle fuel consumption will be required with improvements ranging from ~33- 41 to ~45-61 mpg dependent on vehicle foot print [6].7]. budget to premium. Hence. from 2017 to 2021. ambient temperature and altitude range at the same time as minimising fuel consumption. Going forward. The near zero criteria emissions requirements [2] drive the need for SULEV30 products in the 2020 timeframe where NMOG+NOx tailpipe emissions must be below just 30 mg/mile over the transient FTP75 drive cycle. together with corresponding customer demand for improved fuel economy from automotive manufactures have resulted in new technology and operating strategies being researched and introduced. there has been key strategic direction of powertrain downsizing applied industry wide supported by the introduction of new charged air boosting system technology [8.4]. The fleet average CO2 target for 2015 is 130 g/km CO2 (average vehicle weight). Examples throughout the market from. Discussions are also underway for further legislative actions relating to Real Driving Emissions (RDE). In similar time frames to the CO2 legislation in Europe.9] Examples of downsizing strategies and the introduction of Diesel powertrains with low fuel consumption and high specific power outputs utilising advanced boosting system concepts can be seen industry wide.3. On Diesel powertrain applications.5]. as well as the focus on combustion efficiency and friction reduction. Additionally.In Europe. Further steps to improve the CO2 performance of the Diesel powertrain are currently under investigation with increasing focus on downspeeding [13.6. This presents further challenges to the automotive manufacturer to ensure feedgas emissions and aftertreatment conversion efficiencies are maintained over a wide speed / load. For North American markets the focus on CO2 has also grown with the introduction of challenging Green House Gas (GHG) targets and corresponding Corporate Average Fuel Economy (CAFÉ) legislation [2. legislative bodies have introduced CO2 fleet average targets that challenge automotive manufacturers to introduce new powertrain technology in support of achieving much reduced fuel consumption over the New European Driving Cycle (NEDC) drive cycle. recent focus on legislation relative to criteria and CO2 emissions. These limits are envisaged to become more stringent with the proposed introduction of an EU6 Stage#2 limit forecast to be around 2018 [5]. Given the recent focus in North America on fuel economy there is a growing popularity for Diesel vehicles in the market adding further complications to the automotive manufacturer in meeting stringent criteria emissions limits outlined in LEVIII / Tier3 [6]. Emerging markets in China. the Criteria emissions legislation relating to EU6 Stage#1 for Diesel vehicles from 2014 onwards outlines stringent limits for NOx emissions over the NEDC drive cycle. with a glide path to 2020 of a fleet average CO2 target of 95 g/km CO2.4. further challenges will be introduce with the adaptation of more transient cycles such as those found in the World harmonized Light-duty Test Procedures (WLTP) [2. India and South America have followed the legislative trends of both Europe and North America and hence it can be expected that these markets will also drive automotive manufacturers to implement novel technology to support both low criteria and CO2 emissions. ensuring automotive tailpipe emissions under ‘normal conditions of use’ are regulated [3.14] where two methods of downspeeding are envisaged [9]: 208 . including Toyota [8]. Volvo [10]. Utilising a Jaguar Land Rover Range Rover Evoque vehicle the project aims to develop on the latest Diesel engine architecture.19. 1. In order to support the downspeeding concept of carrying over drive ratios and utilising shift strategy. 1. Further examples of the introduction of compound boost systems utilising both turbochargers and superchargers can also be seen in literature as a strategy for supporting improved transient performance and further enabling downspeeding [9. however this does present further challenges in terms of transient performance of the powertrain. together with advanced active NVH technology to support delivery of a 99 g/km compact premium sports utility vehicle. Moving to downspeeding utilising the gear shift strategy is preferable in maintaining vehicle capabilities.17] as well as the development of electrically driven compressors to provide instantaneous air charge delivery [18. Starting in 2013 the project has progressed with initial analytical assessment of the air charge boost system configuration now completed.10].2 The Provoque Project To support technology transfer Jaguar Land Rover (JLR). 209 . 48V Mild Hybrid Belt Starter Generator (BSG) and e-clutch (manual clutch pedal with electronic rather than mechanical actuation of the clutch).16] with focus on maintaining top end performance. whilst maintaining criteria emissions performance. but also facilitating the low speed transient capability that supports downspeeding. The development of electrically assisted turbochargers offers potential solutions [16. a lightweight low friction cranktrain. 2 ENGINE / BOOST SYSTEM DESIGN 2. Lotus Engineering. Examples of the development of series stage turbocharged boost systems can be seen in literature [15. to support JLR’s development as a stand alone company. charged air boost systems have been the further focus of activities to support a powertrain wide system approach to CO2 reduction.1 JLR Next Generation Diesel Engine As announced in the international press. Provoque project which looks to develop the technology of a 48V electric supercharger and aggressive downspeeding strategy on a Diesel engine within a premium automotive product. a new family of technologically advanced engines have been developed to ensure future vehicle products deliver competitive attributes for the worldwide customer base. Valeo and RaiCam joined as a collaborative consortium to develop 48V electrical technologies and novel clutch by wire ‘e-clutch’ concepts to support the delivery of a premium Sports Utility Vehicle (SUV) that would achieve minimum CO2 through the use of aggressive downspeeding and other technological refinements. 48V electric supercharger. Use of longer gear ratios (internal to transmission or final drive) 2.20.21]. As powertrains become increasingly integrated with electrical architectures the usage of electrical assist to support charged air boosting increases. Carry over drive ratios – downspeed using gear shift strategy Modifying drive ratios does have drawbacks and can result in compromise with respect to other vehicle attributes such as gradability. This paper describes the initial stages of work including analytical investigation relating to the Technology Strategy Board UK Government funded. Valeo 48V Electric Supercharge Figure 1: Computer Aided Design image of JLR Research I4 Diesel Engine 2. The electrical supercharger concept consists of the following sub-systems – centrifugal compressor.000 rpm/s Peak Electrical Powers 7. a 2. rotor assembly.0l four-cylinder Diesel Research engine concept derived from the proposed new engine family was chosen as the base powertrain for the Provoque project. The JLR research engine features an advanced Diesel combustion system utilising the latest high-pressure solenoid injection system.2 Valeo electrical supercharger The Valeo 48V electrical supercharger (VES) concept was first introduced in a 12V variant and has been now further developed into the new 48V hardware for evaluation through the Provoque project.5 kW Compressor Type Radial 210 .0 l Standard Boosting System Single VGT Specific Power ~65 kW/l Specific Torque ~ 200 Nm/l As part of the Provoque project the research engine is planned to be fitted with both a Valeo 48V electric supercharger and Belt Starter Generator device to support attainment of the Provoque project vehicle demonstrator target of 99 g/km CO2.To ensure the research study and technology development described in this publication are aligned to future JLR applications. Table 2: Valeo electrical supercharger specification Parameter Specification Voltage Platform 48V (nominal DC) Motor Three phase Switch reluctance Cooling Air Cooled Peak Speed ~70. Table 1: High Level Details of JLR Research Diesel Engine Engine Parameter Specification General Architecture Inline Four Capacity ~ 2. power electronics and microcontroller. High level details of the research engine are given in Table 1. together with high and low pressure EGR circuits.000 rpm Peak Acceleration ~400. high-speed electric motor. Figure 3: Schematic of the GT Power engine model developed to support assessment of the electric supercharger application to a JLR Diesel engine 211 . A schematic of the one-dimensional model is shown in Figure 3 highlighting the final proposed layout of the innovative boost system configuration. enables controlled demand of the electric supercharger when boost is required. a detailed study of the application of the Valeo 48V electric supercharger was completed. Control of the Valeo 48V electrical supercharger is via CAN and this coupled with an e-hook to the Engine Control Unit (ECU) for the air path control software.High level details of the 48V Valeo electrical supercharger (VES) concept are given in Table 2 with an image of the unit presented in Figure 2. Figure 2: Image of the Valeo electrical supercharger (VES) 3 ONE-DIMENSIONAL MODELLING RESULTS Using a GT Power one-dimensional model of the JLR four cylinder Diesel Research engine. Additionally. The results show a small advantage in transient response with location of the electric supercharger pre the standard turbocharger with approximately a 60ms improvement in reaction time to maximum torque at 1500 rpm. which reduces the compressor performance for the pre-Turbocharger location can also be seen in Figure 4. However. Figure 4: Load Step transient comparison on JLR Research engine investigating position of Valeo 48V electric supercharger Based on the results presented. Results show that the operation of the turbocharger is not significantly impacted by either location. Figure 4 presents results for the electric supercharger located pre (before) and post (after) standard VGT turbocharger. locating the electric supercharger pre the standard VGT turbocharger does impact on compressor outlet temperatures as shown in Figure 4 with an average increase across the speed range of approximately 50°C when compared to the post location. with the standard compressor map supporting both. 212 . the characteristic of the reducing mass flow rate.An assessment of the potential location for the Valeo 48V electric supercharger was completed using the model to determine the optimum location for this supplementary boost device and overall boost system layout. locating the electric supercharge pre Turbocharger does result in higher pressure ratios for the same corrected mass flow rate which will impede the use of low pressure EGR and also there is an impact on the width of compressor map for the performance required. the Provoque project’s direction was to position the Valeo 48V electric supercharger downstream of the standard VGT turbo and develop a new electric supercharger compressor wheel to fully optimise the operation. The corresponding compressor maps for both electric supercharger and standard VGT turbocharger at each location are also shown in Figure 4. Figure 5 shows a comparison of compressor maps for both standard and downspeeded operational cases using two different frame sizes. However. further one-dimensional analysis was completed to understand the benefit of the electric supercharger technology with respect to downspeeding. The data shows that although during the very early phase of load progression the turbocharger matches 213 . Figure 6: One Dimensional simulation of transient load step comparisons on JLR Research I4 Diesel engine with / without Valeo 48V electric supercharger Figure 6 shows the progression of full load torque at six different engine speeds comparing the electric supercharger in combination with the standard VGT turbocharger against the standard VGT turbocharger alone. Figure 5: One Dimensional analysis of Valeo 48V electric supercharger compressor matching Having determined the optimum position for the Valeo 48V electric supercharger and developed an optimised compressor wheel to support the required operational envelope window. further work is required to understand the optimum balance between the electrical energy for the electric supercharger with downspeeding and the electrical energy for torque assist utilising the BSG e-machine. vehicle simulations to estimate the potential fuel economy offered by implementation of the technology and the ability to downspeed were investigated. The results clearly show a benefit and furthermore. Development of the one-dimensional model to include representations of the JLR Diesel Research engine with the both the Valeo electric supercharger and BSG within the chosen vehicle demonstrator platform was completed.5% improvement in CO2 over the NEDC drive cycle. significant improvements in time to full load are achieved. Figure 7 shows simulated engine fuel flow for the baseline (standard VGT & vehicle/transmissions shift schedule) together with the application of the electric supercharger and associated downspeeded operation. the downspeeded operation was characterised by shifting into a higher gear during steady cruise operation. Figure 7: One Dimensional vehicle simulation of JLR Research engine operating with Valeo electric supercharger over an NEDC drive cycle Following evaluation of the location of the electric supercharger. Figure 7 also highlights the potential benefit of torque assist from the application of the Valeo 48V BSG unit in conjunction with downspeeding enabled by the electric supercharger.the electric supercharger (approximately up to 40% full torque). optimisation of the compressor wheel to support both normal and downspeeded operation and transient evaluations of energy consumption. In this instance. Noting that the legislative requirement for the state of charge of the energy storage devices is that it must remain neutral over the cycle. although the standard VGT turbocharger does close the gap providing responses closer to that of the new technology. 214 . As engine speed increases the improvements in time to full load using the electric supercharger are still apparent. Further transient modelling was undertaken with the aim of estimating the energy required by the electric supercharger during various vehicle drive-cycles when utilising a ‘down-speeded’ gear shift strategy. simulations for the Provoque project demonstrator vehicle estimate a 4. “LEVIII and CAFÉ 2025 – Innovative Measures for Compliance of Most Stringent Legilsative Demands”. D. T. Charlton. Joschka Schaub. Sebastian Visser. without any of whom this project would not have been possible. “Shaping the Future – Innovations for Efficient Mobility”. Optimisation of the electric supercharger and standard VGT turbo have been completed using one dimensional analysis and the validation of this analysis is planned during the next stage of the Provoque project. Koji Ishizuka. 2013 3. M. S. Dr. H. Olaf Erik Herrmann. 22nd Aachen Colloquium Automobile and Engine Technology. Dr. Dipl.-Ing. Corning Incorporated. The next steps within the Provoque project will focus on developing the control architectures required to support physical operation of the boosting concept and validation of the fuel economy benefits through vehicle correlation testing. “Combustion Improvement and Emission Control Technologies Supporting the New Cycle Requirements for Passenger Car Diesel Engines”. Nikolas Pörtner.The results presented in this paper highlight the potential of implementing an electric supercharger to support downspeeding thereby improving fuel consumption. Dr. “Next Generation of Common Rail Diesel Injection System Featuring Piezo Injectors with Direct-Driven Needle and Closed-Loop Control”. Greenhouse Gas Emissions”. 2013 2. Marcel Wüst. “Diesel Engine Technologies Enabling Powertrain Optimization to Meet U. Thorsten Schnorbus.-Ing.-Ing. Ansgar Christ.-Ing. REFERENCES 0. J. 2013 6. Dr. Internationales Wiener Motorensymposium. Volkmar Denner. Johnson. Dipl. “Electrified Powertrain at 48 V – More than CO2 and Comfort”. Dirk Queck. Katsuhiko Takeuchi. Alexander Trofimov. Friedrich Kapphan. “Vehicular Emissions in Review” SAE Paper 2013-01-0538. Dr. Schaub. Marc Uhl. Nicolas Nozeran. Dr. which is enabled by the transient response improvement of an electric supercharger. Dr. 2013 215 . Ing.-Ing. Scassa. 34. Stanton. Ken Uchiyama. Dr B. Vajapeyazula. Stefan Lehmann. 4 CONCLUSIONS The early phases of the Provoque project investigation into the application of a 48V electric supercharger have developed the optimum location for the concept in conjunction with a standard single VGT turbocharger and indicated potential fuel saving available with downspeeding.-Ing.-Ing. Dr T. Dipl. Ing. SAE 2013-24-0094. Holderbaum. Nanjundaswamy. Korfer. Dipl. Schnorbus. 22nd Aachen Colloquium Automobile and Engine Technology. 22nd Aachen Colloquium Automobile and Engine Technology. and P. Dipl. 2013 1. Dr D.S. Detlev Schöppe. Dipl. T. 2013 4. Tomazic. 2013 5. 22nd Aachen Colloquium Automobile and Engine Technology. 5 ACKNOWLEDGEMENTS The authors of this paper would like to thank all of the other members of the Provoque project consortium for their involvement and the Technology Strategy Board for their continued support. “Development of High Speed Motor and Inverter for Electric Supercharger”. Dr. N. Tavernier and S. Eriksson. J-E Larsson. Crabb. 2012 17. Lee. Stanton. Hammer. J. 2013 8. 2013 18. S. R. and P. MTZ 09I2013 Volume 74. Martinez-Botas.0-Liter Fuel-Efficient Diesel Engine”. 2013 216 . A. Nishiwaki. SAE 2012-01-0709. E-Boost and Turbocompound Concepts”. K. “In Search of the Optimal Future Powertrain”. T. K. Tanaka. H-J. Langen. Iezawa. Forissier. “New Modular Engine Platform from Volvo”. K. Capon and L. Durrieu. 2013 21. D. C. J. “The Electric Supercharger”. Collings and K. J. Lindell. Chisaki.-H-J. T. “Diesel Engine Technologies Enabling Powertrain Optimization to Meet U. 33. Song and D. Neußer. Leufven and A. 2013 9. “Design and Characterization of an E-booster Driven by an High Speed Brushless DC Motor”. 2013 15. 2012 20. 2012 12. SAE 2012-01-1735. J. Charlton. M. M. T. SAE 2013-24-0094. S. Internationales Wiener Motorensymposium. Engler. Burke. H. Smith. O. L. “Downspeeding a Light Duty Diesel Passenger Car with a Combined Supercharger and Turbocharger Boosting System to Improve Vehicle Drive Cycle Fuel Economy”. Yoshijima. A. P. T. H. Internationales Wiener Motorensymposium. 2012 14. Dorenkamp. Prof. N. “Development of a New 2. R. “Assessing Boost-Assist Options for Turbocharged Engines Using 1-D Engine Simulation and Model Predictive Control”. 2013 13. Equoy. Jauns-Seyfried. Yamada. 34. M. SAE 2013-01-0932. M. D. Thomasson. 2012 16. P. Kikuchi. Morinaka. Jelden. MTZ 5 07-08I2012 Volume 73. Schmidt-Sandte. Zechmair. Brace. 33. Zhang. Vajapeyazula. SAE 2012-01-0713. Glover. G. 2012 11. Dr. Kang. T. K.S. Dr. V. Weber. An. J. Y. SAE 2013-24-0122. SAE 2013-01-0310. Krause. Terdich and R. Akehurst. “Scalable Component- Based Modeling for Optimizing Engines with Supercharging. Darlington. J. S. Oda and K. “Enhancing Power Density with Two- Stage Turbocharger”. P. Q. Internationales Wiener Motorensymposium. SAE 2013-01-1762. Wu. Criddle. Surbled. Goto and B. Manabe. D. Greenhouse Gas Emissions”. SAE 2013-01-0931. M. “Simulation Study of the Series Sequential Turbocharging for Engine Downsizing and Fuel Efficiency”. Cieslar. “Volkswagen’s new modular TDI® generation”. Picron. S. Okamoto. Kahrstedt. S. Wetzel. Somhorst. Dr D. “Future Mobility Solutions of the BMW Group”. 2013 19. Menegazzi. SAE 2013-01-0935. Lee. Fleiss. “Experimental Efficiency Characterization of an Electrically Assisted Turbocharger”. 2013 10.7. D. H. O. Y Hirai. The VGS may. 2014 217 . The Variable Geometry System (VGS) is a validated technology. This paper describes the effects of an EAT mounted on a passenger car diesel engine (D/E) on fuel consumption during the New European Driving Cycle (NEDC) and on engine transient response. An EAT is known as the system which can improve the transient response and Brake specific fuel consumption (BSFC) [3-4]. 1 INTRODUCTION Automotive companies around the world are facing strict emission and fuel economy regulations. Good turbocharger transient performance is necessary. Engine downsizing and downspeeding are important concepts within this process. Japan ABSTRACT An Electrically Assisted Turbocharger (EAT) is one of the effective tools for improving transient response and fuel consumption.Electrically Assisted Turbocharger as an enabling technology for improved fuel economy in New European Driving Cycle operation T Suzuki. lead to increased back pressure and. turbochargers offering a good balance between steady-state and transient performance play an important role in realizing low emission transportation systems. the Real Driving Emissions (RDE) regulation is proposed for Euro 6-2. IHI Corporation has developed simulation tools for turbocharger fuel economy effect estimation during vehicle running. applicable from 2017[1]. while offering low fuel consumption and good control of exhaust gas recirculation (EGR) gas ratio. The EU. This paper _______________________________________ © The author(s) and/or their employer(s). mainly applied on diesel engines (D/E). Downsizing is commonly applied in conjunction with turbocharging to improve environmental performance while maintaining satisfactory drivability. if vehicle transient response after engine downspeeding is to be retained. vehicle performance improvement is required in order to meet such regulations. N Ikeya IHI Corporation. Thus. US and Japan are further required to reduce CO2 emissions by 20% to 45% due to increasing environmental concerns [2]. The validity of the calculation model is confirmed by comparing simulation results against results of transient and steady tests conducted on an engine test bench. able to achieve good transient response at low engine speeds. In the EU. as predicted by means of engine simulation. thus. however. degraded fuel economy. 0L D/E is assumed as the target for EAT study. The engine model validity is confirmed by comparing against experimental steady-state and transient results. which is recovered by engine alternator generation during vehicle deceleration. 2 VALIDATION OF CALCULATION TOOL IHI Turbocharged Engine Simulation (ITES) software is a 0-D engine simulation code used for this study [5]. especially at part load conditions.2L D/E by comparing simulative results with experimental steady-state and transient results from an engine test bench.000 T/C type REV4 (motor turned off) Compressor wheel diameter mm 51 Turbine wheel diameter mm 44. The EAT consumes electrical power during the assist period. 2.2L Cylinder Layout L4 Rating Torque Nm@rpm 340@2. Equipment which is used in this experiment is state-of-the-art and measuring precision is sufficiently high.000 Rating Power kW@rpm
[email protected] Steady State calculation Steady-state comparison is shown in Figure 1. The simulated engine torque. Part load comparison at 3000rpm is shown in Figure 2. enabling agile engineering from model tuning to execution [6-7]. Table 1 Engine specification for validation Engine type Diesel Engine with HP-EGR Displacement 2. 218 . despite the presence of complicated engine systems such as variable air intake and EGR valves. A turbocharger properly sized for a 2.focuses on an EAT mounted on a passenger car D/E and examines the effects thereof on fuel consumption during the New European Driving Cycle (NEDC) and on engine transient response [4]. It is necessary to model both the turbocharger mechanical loss and the engine heat loss when simulating engine bench experiments. Additionally. A comparison of simulation results from a well-tuned model and experimental data is shown as confirmation of the calculation accuracy.5 2.0L D/E with VGS and EAT is assumed herein and an optimum combination of certain operating parameters. minimizing fuel consumption and EAT power consumption is sought. The analysis model validity is confirmed with 2. the turbocharger and engine heat capacity must be considered at transient operation. In the meanwhile. ITES requires reduced computational time compared to commercial 1-D simulation tools and is based on the filling and emptying method. The EAT control logic is designed so as to be a minimum of required modifications for EAT adaptation by vehicles utilizing 12V batteries. The target engine and turbocharger specifications are summarized in Table 1. The EAT potential for fuel consumption reduction during mode operation and for transient response improvement is then presented. EAT is not used in this validation. Simulation results are sufficiently accurate at part load. the results presented are thus within the bounds of this limitation. Specific Fuel Consumption (SFC) and difference between boost and back pressure match well to experimental results under full load conditions. 20Nm 5g/kWh 20kPa 20krpm Figure 1 Comparison of experimental data and calculated data at full load 10kPa 5g/kWh 10krpm 50K Figure 2 Comparison of experimental data and calculated data at part load 219 . 500rpm Figure 3 Experimental condition of transient mode 0.2 Transient calculation A similar procedure is used for model validation under transient operation. which is linked to the exhaust manifold pressure measurement system delay.02kg/s 20krpm 20kPa 20kPa Figure 4 Comparison of experimental and calculated data in transient mode 220 . For this calculation. Some delay of the experimental exhaust manifold pressure is observed. while the comparison between experimental and calculated data in transient mode is shown in Figure 4. turbine speed and intake manifold pressure show same trends as the experimental data.2. The simulated air flow rate. the engine speed and the acceleration pedal opening are used as inputs. Those parameters are presented in Figure 3. 3 EFFECTIVENESS OF EAT In this chapter, it is investigated the EAT impact on an assumptive 2.0L D/E, with a focus on fuel consumption reduction and transient response improvement. 3.1 Calculation condition The specification of the target engine and EAT is summarized in Table 2. The engine utilizes Low Pressure loop EGR (LP-EGR). The EGR ratio is controlled by valves installed at the recirculation route and exhaust pipe. Electrical power regeneration by engine alternator is applied during vehicle deceleration and it should be noted that only power generated in this manner is available for EAT operation. Table 2 Engine specification for EAT study Engine type D/E with LP-EGR Displacement 2.0L Cylinder Layout L4 Rating Torque Nm@rpm 375@1,600 Rating Power kW@rpm 110@4,000 T/C type REV4 Compressor wheel diameter mm 49 Turbine wheel diameter mm 39 Rating power of motor kW/rpm 1.5kW/70,000-128,000rpm 3.2 Effect of EAT on fuel consumption during NEDC Vehicle speed is input as target in this study, while the acceleration pedal opening is under feedback (F/B) control to follow vehicle speed. The driver model is tuned by changing the PID coefficients of F/B control. The VGS control logic, the motor and the generator operation logic of the EAT, as well as the gear set of the transmission are set as parameters for which an optimum combination minimizing fuel consumption and EAT power consumption is sought. VGS nozzle control EAT motor control Switch 1 Switch 1 Switch 2 Switch 2 Figure 5 Example of control map of the VGS nozzle VGS nozzle and EAT motor are controlled as shown in Figure 5. The VGS nozzle and EAT motor are under F/B control with boost pressure as target above and under open loop control below the switch line. Two switch lines (Switch 1 and Switch 2) were prepared. With respect to EAT motor control, three maps for F/B control are used: Powerful operation, Proportional operation and Economy operation. The difference between these maps is the maximum assist power of EAT motor. Powerful operation assists in a proactive way, while economy operation assists in a power-saving manner. There are also two modes of EAT motor open loop control: control with motor assistance and control without assistance (motor off). The assist is always off at 5th and 6th gear and during shift-up operation. Two patterns of transmission gear set (T/M) are prepared: Gear Set 1 as the base gear set and Gear Set 2 for downspeeding. 221 The gear ratio of Gear Set 2 is smaller than the base one. The down-speed ratio at 1st gear is smaller than other gear position to avoid start ability degradation, while downspeeding is not applied at 6th gear, so as to maintain vehicle top speed. The gear sets are summarized in Table 3. Table 3 Downspeeding Ratios Down-speed ratio from Gear Set 1 % Remark Gear 1 2 3 4 5 6 Gear Set 2 8.5 10.7 8.7 4.6 4.1 0 Down-speed version The NEDC simulation is conducted with the 8 cases indicated by Table 4. In CASE2, CASE3, CASE4 and CASE5, the motor assist turns on only while the vehicle is accelerating; in CASE6, CASE7 and CASE8, the motor assist turns on not only when accelerating but also during constant speed running, as later described. Fuel consumption during NEDC is summarized in Figure 6. Table 4 Calculation pattern for NEDC study CASE VGS EAT Motor Motor Motor T/M Shift-up switch Switch Open F/B Turn- gear timing loop off set Set 1 (Base) SW2 - - - - 1 1 2 SW1 SW2 Off Pro X 2 1 3 SW1 SW1 Map Pro X 2 1 4 SW1 SW2 Off Eco X 2 1 5 SW1 SW2 Off Pow X 2 1 6 SW1 SW1 Map Pro - 2 1 7 SW2 SW1 Map Pro - 2 1 8 SW2 SW1 Map Pro - 2 2 SW1: Switch 1/ SW2: Switch 2 Map: Map control/ Off: motor turn off Pow: Powerful operation for motor Pro: Proportional operation for motor Eco: Economical operation for motor Shift up timing set 1: 1700rpm Shift up timing set 2: The engine speed to keep the same vehicle speed as gear set1 Motor turn off: Motor turns off when current boost pressure is over target 2.2% Figure 6 Fuel consumption during NEDC 222 In CASE2, CASE4 and CASE5, the motor assist turns on while the vehicle is accelerating during mode. Among these, CASE4 (Eco operation) shows highest electrical power consumption and maximum fuel reduction. This is due to the EAT turn off switch, which is the function turning the EAT off when current boost pressure reaches target boost pressure. The motor only resumes the assist when the deviation reaches 5%. In case of abrupt acceleration, operations Pow and Pro can offer more electrical power than Eco, which can lead to the motor turning off caused by boost pressure reaching the target value. Despite Eco operation assisting with less power, it continues to assist for longer because target boost is not achieved. As a result, Eco operation consumes the most amount of electrical power, which leads to the highest fuel reduction rate. The assist power difference between Pow and Pro operation is little because NEDC has a high operating proportion at loads below 50%. This means that the Pro operation would be sufficient for low load modes such as NEDC. Therefore, CASE3 and CASE6, where the motor assist turns on not only while accelerating but also during constant speed running, both use Pro operation. A comparison between operation with and without turn off switch is conducted, but this time avoiding motor turn off due to target boost pressure obtainment and, thus, allowing the motor to operate for a longer time period. As can be seen, CASE6 exhibited lower fuel consumption, which is the reason it was qualified for comparison against CASE7, The difference between these two is VGS nozzle controlling, as is shown in Table 4. CASE7 is shown to have improved fuel consumption. This is because the F/B area of Switch 2 is broader than that of Switch 1. Thus, F/B control uses nozzle openings of higher efficiency. With a suitable open loop control map, the difference between the two will reduce. Also, a wider feedback area is better in light of performance deterioration due to age. Eventually, CASE7, which exhibits best fuel economy of all operating scenarios is qualified for further downspeeding and renamed as CASE8. As expected, CASE8 shows the overall best fuel economy, achieving 2.2% decrease of fuel consumption, with 24.8kJ used for the EAT, which corresponds to 0.1% of total fuel energy. Figure 7 shows a number of parameters from CASE1 and CASE8 between 800s and 850s during NEDC operation. The fuel consumption reduction of CASE8 is caused by a pumping-loss decrease and engine down-speed. The degradation of EGR ratio cannot be observed herein. This study reveals that the EAT motor can reduce the fuel consumption by motor assist not only during vehicle acceleration, but also during constant speed running; however, motor operation should be considered well. As mentioned previously, a 12V battery vehicle is assumed. Fuel economy can further improve by increasing the assist power. 223 Figure 7 Detail behavior of each parameter 224 3.3 Effect of EAT on full acceleration A comparison of vehicle acceleration performance is conducted next. 3 cases are used: no assist (base), EAT assist and EAT assist with downspeeding (Table 5). The initial condition is vehicle constant running at 2nd and 3rd gear. The acceleration performance is assessed as average acceleration to target vehicle speed. The initial speed and target speed are from 17.5km/h to 40km/h at 2nd gear and from 27.5km/h to 60km/h at 3rd gear. The calculation results are shown in Figure 8. Table 5 Calculation condition of full acceleration CASE Gear set Motor assist Remark 1 1 - Base 2 1 X EAT assist 3 2 X EAT Assist + downspeeding 2nd gear 3rd gear Figure 8 Comparison of acceleration For CASE2, it is found that the acceleration performance is improved by 12.5% at 2nd gear and by 8.5% at 3rd gear. In CASE3, acceleration performance is improved by 4.9% at 2nd gear but remained equal to base scenario at 3rd gear. 225 4 SUMMARY AND CONCLUSIONS This study attempts to predict the effects of an EAT. Model validation is initially conducted by comparing simulative results with experimental data from an engine test bench, both in the steady-state and transient regime. An examination of EAT impact on NEDC fuel consumption and on transient response follows: - An EAT can reduce fuel consumption by 2.2% during NEDC, assuming optimal EAT motor and VGS operation (CASE8 in table4). - An EAT can improve vehicle acceleration by 12.5% at 2nd and 8.5% at 3rd gear without downspeeding (CASE2 in table 5) and by 4.9% at 2nd gear with downspeeding (CASE3 in table 5), while maintaining the base acceleration at 3rd gear. These conditions below are necessary to achieve the performance improvement shown above: - Vehicle with the engine and the transmission designed to be downspeeded. - The combination of VGS nozzle control by feedback in a wide engine operating condition with mild electrically assist in low engine load condition. - Full power electrically assist when vehicles accelerate with engine full load. To conclude, the EAT can simultaneously improve fuel economy and transient performance, provided that the EAT is applied at a downspeeded vehicle. A vehicle with 12V battery is assumed herein and EAT operates under this limitation. The result will, however, change depending on the battery system of the vehicle. 5 REFERENCES [1] Shigeo Furuno. (2013) Global Trends of Emission Regulation and Powertrain Technologies for Vehicles, Journal of Society of Automotive Engineering of Japan. [2] IEA (2012a), Technology Roadmap: Fuel Economy of Road Vehicles, IEA, http://www.iea.org/ [3] Nicola Terdich, Ricardo F Martinez-Botas, Alessandro Romagnoli, Apostolos Pesiridis (2013) Mild Hybridization via Electrification of the Air System: Electrically Assisted and Variable Geometry Turbocharging Impact on an Off-road Diesel Engine. ASME Turbo EXPO 2013. [4] IHI Corporation (2011) Electrically Assisted Turbocharger the power electronics which improve the good fuel economy. IHI technical Review, vol.51, no.1, pp.14-15. [5] N Ikeya, H Yamaguchi, K Mitsubori and N Kondoh. (1992) Development of advanced model of turbocharger for automotive engines. SAE paper 920047. [6] Watson N and Janota M.S (1982) Turbocharging the Internal Combustion Engine. The MacMillan Press, Ltd. [7] Georgios Iosifidis, Jason Walkingshaw, Bernhard Dreher, Dietmar Filsinger, Nobuyuki Ikeya and Jan Ehrhard. (2013) Tailor-made mixed flow turbocharger turbines for best steady state and transient engine performance. 1st International Conference on Engine Processes. 226 Development of a high-efficiency commercial-diesel turbocharger suited to post Euro VI emissions and fuel economy legislation J Watson, R Lotz, D Grabowska, J Moscetti, T House, S Scott, R Vemula BorgWarner Turbo Systems, USA ABSTRACT The development period post Euro VI / EPA ’10 finds diesel engine OEM’s seeking combined combustion and aftertreatment solutions to the challenges posed by new fuel economy-focused legislation regulating diesels. It is the intent of this paper to discuss efforts underway, and results achieved to date, aimed at development of a high-efficiency commercial vehicle turbocharger supporting these fuel economy aims. This new high efficiency turbocharger is projected to achieve in excess of 60% overall efficiency and thus should support well any engine- or vehicle-level effort targeting attainment of the new aggressive fuel economy standards. NOMENCLATURE BTE – brake thermal efficiency for h◦ – change in enthalpy across the the engine stage SCR – selective catalytic reduction Q – volume flow through the stage EGR – exhaust gas recirculation – rotational velocity of the stage NOx – oxides of Nitrogen including – efficiency of the stage, reported NO and NO2 here on a t-s basis PM – particulate matter P – pressure Ns – specific speed of the stage M – mass flow Ds – specific diameter of the stage T – temperature D – physical diameter of the stage CSLA – constant speed load acceptance 1 INTRODUCTION / MOTIVATING FACTORS 1.1 Regulatory environment External regulatory factors are largely responsible for driving a careful assessment of approaches to commercial vehicle engine programs targeted for production in the 2017 and forward time frame. Pending governmental regulation in the US and European Union mandates aggressive reductions in fuel consumption, and consequently in CO2 emissions, with allowable emission of the traditional diesel pollutants oxides of nitrogen (NOx) and fine particulates (PM) effectively holding flat, at least for now. While the collective body of national and state regulators may consider further reductions in allowable NOx near term (additional reductions in NOx of up to 75% are possible beyond EPA ’10), the macroeconomic motivators of (national) desire for energy independence and (owner operator) desire to minimize fuel costs coupled with EPA’s newfound emphasis on fuel economy are expected to cement fuel economy at top of the diesel agenda near term. _______________________________________ © The author(s) and/or their employer(s), 2014 227 1.2 OE response – directions in diesel engine development The fuel economy benefits of engine downsizing, the act of reducing engine displacement to effect a reduction in friction (enabled by use of a turbocharger or supercharger to augment engine brake power), and engine downspeeding, the act of using transmission and related vehicle level changes to narrow the effective speed range of the engine for purpose of reducing friction and enhancing combustion efficiencies, have been explored for some time now and their implementation is expected to accelerate with the pending legislation (1). Similarly, use of Miller Cycle to modify the effective in-cylinder charge temperature with positive consequences for NOx production (while enabling fuel economy friendly injection timing changes) is also viewed as a viable technology for post Euro VI / EPA ’10 operation. Accompanying the potential introduction of the above strategies is a general trend towards universal acceptance of SCR as the industry favored technology for managing tailpipe NOx emissions against Euro VI and EPA ’10 mandates. The relative success of SCR as a technologically effective and robust tool for reducing NOx emissions is viewed as an enabler for the next round of fuel economy enhancements, if for no other reason than it tends to better support more fuel- economy friendly positive engine scavenging. Any new air handling system design targeting this type of application must consider these facts in order to provide best fit to the engine. 1.3 Air system architecture selection It is well established that the targeted performance and emissions outputs of the diesel engine (as prescribed by requirements of torque, air to fuel ratio, fuel consumption, and rates of EGR) drive specification of the air system employed. Elimination of or reduction in the usage of high pressure loop EGR, as enabled by the relative success of SCR, tends to have a simplifying effect on the air system. Namely it: Moderates the amount of boost air required, suggesting single-stage turbochargers where two-stage units might have been required prior, Relieves the turbine stage from the need to drive high pressure loop EGR, allowing for a reduction in the associated engine pumping work, Leads to selection of turbine stages with higher swallowing capacity (and also likely better efficiency), If coupled with engine downspeeding as a strategy for friction reduction, drives a reduction in required compressor and turbine operating ranges. Technologies such as vane diffuser for compressor and fixed nozzle ring for turbine thus suggest themselves for inclusion in the turbocharger as they offer the promise of enhanced operating efficiency (though usually over a limited range of operation), Reduces the necessity for variable geometry turbocharging for many applications. Elimination of the VTG mechanism simplifies the turbocharger, offering in the trade a more robust hardware solution and one that should ultimately facilitate attainment of best overall turbocharger efficiency. The turbocharger suggested by the above is a single stage unit incorporating stages capable of moderate to high speeds and pressure ratios, running fixed geometry for turbine, and delivering best possible efficiency in the steady state. 228 has considered optimization on multiple levels – at the impeller. System level optimization requires optimization of individual components as well as knowledge of how those components function together as a system. the turbine stage has been optimized using an empirical.4 Other considerations Beyond top level architecture. experienced-based approach. As such the primary focus of the development is maximization of the turbocharger overall efficiency while concurrently delivering on the requisite flow. but not to excess so as to adversely impact performance? Match – i. selection not only of the preferred blading for each of the compressor and turbine stages. 1. and for the turbocharger as a whole. and package requirements normally associated with commercial vehicle applications.e. driving increased emphasis on delivery of improved stage efficiencies at part load operating conditions. The uniqueness embodied by the development lies in the holistic. The turbocharger development effort described herein.3. Performance maps for the baseline stages have been obtained and are used to reference suitability of both flow capacity and pressure ratio for the developed stages.5 The development approach: what have we done? This paper describes a proactive turbocharger development program meant to address the aspiration needs of a diesel engine for commercial vehicle targeting compliance with post Euro VI / EPA ’10 emissions and fuel economy regulations. Efficiency targets for the developed stages have been set in relation to the peak values shown on the baseline maps. The baseline turbocharger is constituted by a single stage employing a fixed geometry wastegated turbine housing.e. and targeting the mission described above in sections 1. and in the present case this means traditional high specific speed turbocharger stages flowing the correct ranges for the engine application targeted. but also selection of the proper sizes of each in relation to both the anticipated operating ranges and in relation to one another. but consideration has also been given to the level of efficiency increase representing a true step change when compared to today’s aero product line. i. the following must be considered in defining the preferred turbocharger for the mission: Turbocharger mechanical elements (as they support the overall performance proposition of the turbocharger). We have here attempted to employ multiple approaches to optimization of performance: the compressor performance has been optimized using both empirical and state of the art numerical optimization approaches. speed. at the stage. It will be suggested herein that a thoughtful turbocharger match considering each of the above points. informed not only by consideration of the performance and emissions objectives of the engine but also by a thorough knowledge of the performance sensitivities of the individual stages. durability. The first step has been to identify a baseline ‘off the shelf’ turbocharger best aligned with the mission targets.1.1-1. bearing and thrust components – are ball bearings preferred over more traditional fluid film journal bearings? How are thrust loads best managed such that adequate thrust capacity is provided. will be essential to attainment of best overall air systems performance. system level approach taken to ensure best overall performance. The importance of fuel economy over a full drive cycle. The performance of the turbocharger thrust components has been considered via CAE and geometrical 229 . it should be noted that the impeller design direction to be followed here involves a general increase in rotor group inertia as described below. defined according to Ds D h 1 4 Q inflow angle – i.g. With volume flow and enthalpy change across the stage fixed as constraints. By making appropriate reference to the Balje diagram early in the design process. For fixed efficiency the trend in size with decreasing specific speed should be apparent.6 Design direction for stages It is common industry practice to employ non-dimensional parameters (e. both with the baseline turbocharger and with the developed turbocharger.1 to 1 in the units shown. trades of efficiency v. In this way it is believed the best overall system level performance has been realized. Reynolds number or Mach number) for classification purposes. a change to one of specific speed or specific diameter or both is required. a Balje diagram can be used to relate single stage turbomachines categorized by specific speed. Our applications will constrain consideration of specific speeds to the range 0. Lastly. specific diameter. axial. Selecting the appropriate non-dimensional parameters allows relation of candidate designs to historical efficiency trends (5). the equations defining specific speed and diameter may be used to compute new speeds and diameters for the candidate designs. Ball bearings have been considered based on documented and clear evidence of their capacity to improve performance of the overall turbocharger at low speeds. both relative to one another and to known regions of high operating efficiency (2). package size can be inferred at the outset and appropriate decisions taken regarding preferred bounds of the design space. In particular. or mixed. It will be the intent of this paper to describe approaches applied and outcomes observed at each level of development for this turbocharger.changes subsequently have been made to drive best combined turbine efficiency.e. A Balje diagram for turbines and expanders is depicted in Figure 1. defined according to Q Ns h 3 4 specific diameter (2). These impacts have been considered in relative terms by using engine cycle simulation tools to assess transient engine response over defined events. and as such is a useful tool for comparing potential designs. The design geometry providing the best balance of efficiency and size (package and rotating inertia) is usually selected for additional detailed work. Furthermore. the stages have been carefully size-matched based on directed test outcomes suggesting the preferred compressor loading for the turbine developed. In order to improve upon the efficiency of a baseline design. 1. as one of radial. The Balje diagram presents the following non-dimensional parameters: specific speed (2). 230 . inflow angle and efficiency. The numerical optimization employed a hybrid genetic (7) and sequential quadratic programming (8) algorithm to vary impeller hub and shroud contours as well blade angles to build a Pareto front of designs meeting a 231 . Regions of constant efficiency are plotted as contours. and more expensive 3-D CFD methods. and to be manufacturable at competitive cost.4). low specific speed designs have been applied in an attempt to maximize operating efficiencies. The design process is further constrained by needs to maintain structural integrity over a long operating life. and it is understood that any performance uptick so realized will come at cost to rotating inertia of the stages. 2 COMPRESSOR STAGE DESIGN Radial compressors are used almost exclusively in turbochargers due to their compact design. Further refinement is then accomplished with streamline curvature methods that allow determination of blade loading. A separate approach using numerical optimization techniques in combination with 3-D CFD was also employed. and to packaging. to present low inertia for good transient response. Figure 1 Ns-Ds diagram for turbines and expanders. and the designs must be tested to ensure that prediction aligns with tested outcomes. high specific speed designs. wide operating range and low manufacturing cost (4). from Balje (2). Turbochargers historically applied to commercial vehicle applications involve radial flow. as well as preliminary values for inducer and exducer blade angles and area ratios. the preliminary design of the compressor impeller was based on empirical correlation (a Balje diagram for compressor) and simple 1-D analysis tools (3. All of these methods suffer to some degree from not being able to represent reality faithfully. In the present study radial flow. For this development. A good design combines efficiency at the target operating point (typically engine peak power) with acceptable surge margin at low rpm engine conditions (peak torque) and altitude margin (ability to deliver acceptable performance at higher pressure ratios and rotation speeds). This set the overall dimensions and speeds. e. Structural integrity for low cycle fatigue required the impeller to have peak stresses due to centrifugal loading below a specific threshold at operating speed. It is believed that the map topology delivered supports well the aim of improved fuel economy over targeted commercial vehicle drive cycles by improving upon stage efficiency in regions of the map often important for part load operation. Once the aerodynamic geometry was frozen. and only subsequently to optimize the structure for both high and low cycle fatigue loadings. a numerical optimization scheme using a combination of a multi- objective genetic and sequential quadratic programming algorithms was used to build a three dimensional Pareto front characterizing the trade-off between the three objectives. fillet radii and backwall shape were seen to be the dominant variables influencing stress levels. and pressure ratio. 2. at lower speeds and at higher mass flows. This allowed the aerodynamicist more freedom in exploring different aerodynamic concepts.1 Structural optimization methodology A conscious decision was made to define the aerodynamic shape of the impeller first. Typically 150 to 200 design evaluations were required to build up a Pareto front of sufficient density to pick an optimum. 2.multi operating condition and multi objective design problem. ultimately. Due to the relatively small inducer required by a low specific speed design and the resulting shorter unsupported blade length. i. With four to six variables to describe the backwall shape and fillets and three stress objectives for back wall. For this development impellers had to meet basic structural integrity requirements in order to be stand tested. the optimization problem becomes too complicated to be solved efficiently using a DOE approach. 232 . and so additional work will be considered to improve the map in this region. Larger fillets and a thicker backwall typically reduce stress in these components but also increase the bore stress. The optimum configuration presents a balance of the lowest possible stress in all regions. fillet and bore stress. It is also clear that peak stage efficiency is located away from full-load rated-power operation for applications requiring more aggressive intake manifold pressures. The ‘best’ such result achieved is as depicted in Figure 2. a vaneless design was selected for inclusion in the high efficiency turbocharger due to the attractive combination of stage efficiency. map width. An additional requirement for the impeller to be flank millable for cost effective manufacture required a constant radius fillet at the base of the blades.2 Results achieved – compressor performance For this development both vaned and vaneless diffuser designs were executed and. but required somewhat greater effort to structurally qualify the resulting designs. the result being designs with up to 10% lower stress than those based on standard dimensioning guidelines. high cycle fatigue did not represent a particular challenge for the compressor and could be accommodated with minimal impact to aerodynamic performance. Rather. Three phases of stand test were conducted where candidates of each approach were evaluated. 233 . From there. This design philosophy yields a turbine with high through-flow.09 and a specific diameter of 1.2). The flow parameter was 0. turbine volutes are frequently wrapped or folded away from the center body to accommodate the high flow without increasing diameter.4.85 at expansion ratio 2. Once the preferred 1-D design was settled. Optimized Design Baseline Design Figure 2 Tested performance of a vaneless compressor stage developed for this program.0) (4).8 to 2.0 to 1. frequently operating at sonic velocities and low specific diameter (1. NC USA. Compressor data is corrected to 298K and 1000 mbar. Efficiencies are normalized to those of the ‘off the shelf’ reference stage defining the flow capacity range. typical commercial vehicle operation drives temperatures of the incoming exhaust gas in excess of 750°C and speeds in excess of 500 m/sec for the turbine and the baseline turbine considered here supports both. Data is measurement on the BorgWarner test stands in Arden.04 kg-√K / sec-kPa. Further. Having located the baseline turbine design on the Ns-Ds diagram. turbocharger turbines have been designed to a high specific speed (1. In keeping with the small space claim. a one dimensional impeller design using the method described by Whitfield and Baines was performed (5). a fully 3-D design of the turbine impeller was articulated (with a variety of blade angle progressions created). new values of specific speed and specific diameter moving the design in the direction of increasing efficiency were selected. 3 TURBINE STAGE DESIGN Historically.5. As indicated prior in 1. low inertia and small space claim at the expense of efficiency. The baseline stage performance compares well with published results (4). and the designs characterized using tools of CFD. a more traditional approach was used in designing the turbine stage with preliminary sizing supported by the specific speed-specific diameter charts published by Balje (2). The turbine stage selected as baseline for this development had a specific speed of 1. Figure 3 Performance of the low-specific speed turbine developed for this program.2 Results achieved – turbine performance As shown in Figure 3. Based on test results obtained. The case with no fixed nozzle ring. The larger diameter impeller combined with the symmetric volute increased stage diameter by 14% and inertia by 59% above the baseline high specific speed design.96 at a specific diameter of 2. 234 . In the event the developed stage design did not achieve target efficiency. Additionally. The design presents a specific speed of 0. NC USA. Turbine data is corrected to 923K and 1000 mbar. As with the compressor. 3. an iterative process is employed whereby thermodynamically attractive designs are analyzed in FEA with subsequent adjustments made to the design in order to ensure viability of the impeller.1 Structural optimization methodology The turbine impeller is subject to thermo-mechanical constraints which must be accommodated in the design process. 3. 5 were selected for stand test and were also qualified for stress and frequency prior to prototyping and test. an accompanying symmetric turbine volute was designed to reduce stator losses. the 3 best stage designs were used as seeds for the next design iteration. The addition of a fixed nozzle ring has been reported to decrement performance in some cases (6) and increment performance in others. it was surmised that a nozzle ring could potentially boost stage efficiency closer to target and so was considered for inclusion in the high efficiency turbocharger developed here. the final developed turbine stage achieved 109% of the baseline combined turbine efficiency at 2.Of the 11 candidate impeller designs conceived.17.2 expansion ratio. Data is measurement on the BorgWarner test stands in Arden. Efficiencies are normalized to the peak for the baseline stage. A total of 3 design iterations were completed. The process is repeated until both performance and durability goals are achieved concurrently. The most complex allowed optimization of the vane angles. This prediction was then compared to measured performance data from test stand. arc length. For this turbocharger tapered-land thrust bearings are considered. A sensitivity analysis involving the geometrical parameters of the bearing system was then performed in order to determine efficiency improvements attainable. it must be understood they are achievable via reduction in thrust capacity and so will be difficult to realize for demanding engine applications.1 Optimization of turbocharger thrust components An effort was made to quantify the potential turbocharger level efficiency gains achievable by modifying the stock thrust component geometries. optimization is required to properly balance the competing requirements of maximizing load capacity while minimizing friction power loss. CFD predicted a peak efficiency of 110% of the baseline combined turbine efficiency for the moderately complex nozzle ring with a flow coefficient slightly below target. intelligent trading of excess thrust capacity for performance gain. balanced by an associated reduction in load carrying capability. The final design selected was a more traditional. While these results were deemed encouraging. and radius as shown in Figure 4. The first four were the result of computer optimizations of increasing complexity. 235 . thickness profile and number of vanes. Experimental results indicated a gain of up to 3% in measured combined turbine efficiency. and because the geometrical parameters are easily modifiable in a controlled laboratory setting. As the thrust components themselves tend to impact mechanical efficiency of the turbocharger to great degree. Understanding that mechanical loss dwindles and bearing system loading increases with operating speed. Ultimately the lack of significant performance uptick from the case with no nozzle ring and concern over increased high cycle fatigue risk led to the omission of the fixed nozzle ring from the high efficiency turbocharger configuration. The data is of interest in that it supports improved understanding of the impact of thrust capacity on overall turbocharger performance and further supports. The least complex design modified only inlet and discharge angles and radii. Arc Length Arc Length Radius Height Figure 4 Tapered-land thrust bearing geometry optimized as part of the present study. In the present study specification of thermodynamic and lubricant conditions supported a fluid-film bearing system simulation used to predict the mechanical friction. experience based design. Test results suggested little difference in performance between the stage with moderately complex nozzle ring and the stage with the experience-based nozzle ring design. for select applications.Five nozzle rings were designed for this study. radii. with the running surface and associating annular bearing volume described by a designated height. a decision was taken to evaluate these potentials before moving on to consideration of more obvious alternative technology bearing sets. 4 TURBOCHARGER LEVEL CONSIDERATIONS 4. Typical results obtained by BorgWarner are as shown in Figure 5.Understanding that the typical operations profile for this turbocharger suggests higher speed operation. it is also surmised that there is greater benefit at low expansion ratios because the thrust load is generally highest there. and having satisfactorily quantified the potentials achievable via modified traditional thrust component geometry. Data is measurement on the BorgWarner test stands in Arden. full attention turned to consideration of ball bearings. For each constant speed line. 4. Greater benefit is seen at low speeds because the power consumed by the bearings is a greater percentage of the total power of the turbocharger at lower speeds. helping to overcome the increased inertia of the new impeller designs. Given the obvious benefit to the present development ball bearings have been slated for introduction to the high efficiency turbocharger. This results in higher power consumption by the hydrodynamic thrust bearing whereas the ball bearing cartridge power consumption is not as sensitive to thrust load. Measurement with Ball Bearing Measurement with Fluid-film Bearing Figure 5 Impact of ball bearing on combined turbine efficiency.2 Ball bearings for improved mechanical efficiency Ball bearings have been previously demonstrated by multiple investigators to provide mechanical (and consequently combined turbine) efficiency benefits for turbochargers. 236 . NC USA. Efficiencies are normalized to the peak combined efficiency attainable by running the same turbine aero with journal bearings. This reduction in bearing power consumption is expected to produce better transient response. sensitivity studies were conducted via stand test to quantify the performance variability of each with trim (ratio of inducer diameter to exducer diameter for compressor. the developed stages may be best suited for genset.e. and so these same tests help dimension those effects as well. such simulations help steer the impeller design process by suggesting. for a given engine application. as- designed rotor inertias and engine response again checked for suitability to application. Time to target engine torque. 4. the engine cycle simulations allow the development team to quantify differences in anticipated engine response whereby the off-the-shelf stages are contrasted with the developed stages. For this reason compressor stages delivering a reasonable rotating inertia but wider performance maps better 237 . both constant speed load acceptance (CSLA) and acceleration events were simulated for engine models calibrated to recent vintage (i. In other words. Effectively. engine cycle simulations have been conducted. for turbine. 2010 EPA or equivalent) emissions regulation. Quantitative understanding of both trends specific to the components in question supports optimization of overall turbocharger performance. Performed early in the development cycle. and so the turbine stage developed here was tested with a range of compressor blades and sizes and the impacts to turbine combined efficiency noted. but for a good many commercial vehicle applications a delay of this magnitude will not be significant. and time to target intake manifold pressure are noted in each case. it is generally understood that combined turbine efficiency is a strong function of compressor loading via the dependence of turbine efficiency on blade speed ratio. marine or similar applications where response times are not paramount. Engine response is shown in these studies to be a strong function of compressor map topology. Any deficiency in engine response then motivates consideration of inertia reduction options (design or materials) for the developed stages. It is also understood that the size matching employed impacts thrust loading of the turbocharger. For example. but the stages will likely also have utility for select over-the-road truck or row-crop tractor applications as well. The following are noted: All else being equal. inverse for turbine). housing sizing and. how much steady map efficiency is required to offset a particular increase in rotor group inertia. sensitivity to compressor loading. time to target engine speed. Later in the development cycle actual achieved map efficiencies can be paired with known. aerodynamic tip width of the impeller blades. For some response-sensitive applications this could represent unacceptable delay. For each hardware configuration considered.4.3 Matching – assembling the pieces For each stage developed. which again impacts turbine combined efficiency. not just peak compressor map efficiency (and also not just the stage inertias). but the actual difference in engine response times provided by the lower inertia (and lower efficiency) off-the-shelf stages and by the higher inertia (but higher efficiency) developed stages is predicted to be at most on the order of a few seconds for the engines and events simulated. rotor groups of lower inertia will respond faster.4 Assessing the stage efficiency / inertia trades In order to assess the engine response impacts of the increase in developed impeller inertias. Data is measurement on the BorgWarner test stands in Arden. Compressor data is corrected to 298K and 1000 mbar. Ultimately. Ball bearing use is expected to enhance the tested data as shown. and it is expected that the final outcome will improve. It is of note that this result does not (yet) include the contribution of ball bearings to overall performance. turbine data is corrected to 923K and 1000 mbar. for example a vaneless compressor configuration with aluminum impeller. Projection with Ball Bearing Figure 6 Overall efficiency for a high efficiency turbocharger utilizing low- specific speed stages. based on prior experience. The final value achieved for the high efficiency turbocharger employing low specific speed aerodynamics and running traditional fluid film bearings is 62. will tend to fare better in the transient response simulations than will vaned.0% overall efficiency as supported by the test data presented in Figure 6. this program set as a developmental goal attainment of overall turbocharger efficiency in excess of 60% as measured on test stand. another 2-3 efficiency points at low speed and expansion ratio (the full benefit apparent from Figure 5 is reduced here due to the 238 . narrow map counterparts. but also to motivate consideration of the proper material solutions for each impeller developed. NC USA. and the decision has been taken here to proceed with further consideration of material alternatives to Inconel 713C for the high efficiency turbine impeller where intended for response-sensitive applications. results of these simulations are here meant to assess not just suitability of the developed thermodynamics. 5 RESULTS FOR FULL TURBOCHARGER In consideration of all of the above. encompassing the range of operating conditions simulated. . which is of course less than 1. but also increased inertia and package size. 2007. Baines. April 2002 (8) Adaptive Filter SQP. Inc.multiplication of the combined turbine efficiency and the compressor isentropic efficiency. Longman Group.1. ISBN 0-933283-14-8. IEEE Transactions on Evolutionary Computation. 6. It has been shown that low-specific speed aerodynamics can tend to higher individual stage efficiencies. 2010. Finally. presents opportunity for application of high efficiency low-specific speed aerodynamics. Inc. Design of Radial Turbomachines.. K. Stand results for the high efficiency turbocharger with ball bearings are pending introduction of ball bearing hardware in the proper frame size.Heavy Duty Diesel Engines. O. and Theory. Turco. 1981 (3) Japikse. coupled with the potential to narrow the range of operating speeds via the vehicle-level strategy of engine downspeeding. A Direct Performance Comparison of Vaned and Vaneless Stators for Radial Turbines. when compared to typical high specific speed radial flow solutions. Technology Roadmaps . RD. ISBN 0-933283-03-2. potential improvements to the delivered compressor map topology will be considered. and the authors believe it is clear that relaxation of the negative pressure scavenging requirement associated with the need to drive large amounts of high-pressure loop EGR. The proper air handling hardware will be required to support attainment of those goals. it is the authors’ contention that a carefully-matched combination of low specific speed stages can result in attainment of aggressive overall turbocharger efficiency goals.. Journal of Turbomachinery.. Baines. No. Concepts ETI. 2. London. 1996 (4) Baines. New York. Lecture Notes in Computer Science Volume 6073. Fundamentals of Turbocharging. C. Deb..12/111701. 2005 (5) Whitfield... Vol. Additionally. Turbomachines: A Guide To Design.0). 2012 (2) Balje. Selection. Next steps for the development involve incorporation of ball bearings into the high efficiency turbocharger and also demonstration of stage and solution scalability. D. A. A. John Wiley & Sons. Ricardo plc. p53-61 (7) A Fast and Elitist Multiobjective Genetic Algorithm: NSGA-II. and are not available by time of final draft in February 2014. 978-3-642-13799-0 239 . pp 68-81.. 1990 (6) Spence et al. REFERENCE LIST (1) Such. N. Centrifugal Compressor Design and Performance. 6 SUMMARY / NEXT STEPS The development period post Euro VI / EPA ‘10 finds diesel engine OEM’s seeking combined combustion and aftertreatment solutions to the challenges posed by new fuel economy-focused legislation regulating diesels. N.. Concepts ETI. in this case in excess of 60% overall turbocharger efficiency. Vol 129. as a way to satisfy future HDD engine requirements without the cost and complexity of two stage systems. UK ABSTRACT Future Heavy Duty Diesel (HDD) engines will increase their specific power output. which in turn will increase the overall pressure ratio requirements of the engine boost system. The opportunity for a new. DEFINITIONS AND ABBREVIATIONS BSFC –Brake Specific Fuel Consumption BTE – Brake Thermal Efficiency CARB – California Air Resource Board CO2 – Carbon Dioxide DPF – Diesel Particulate Filter EFTA – European Free Trade Association EGR – Exhaust Gas Recirculation EPA – Environmental Protection Agency (US) EU – European Union GHG – Green House Gases GVW – Gross Vehicle Weight HDD – Heavy Duty Diesel (~2L/cyl engine capacity) MDD – Medium Duty Diesel (~1l/cyl engine capacity) NOx – Nitrogen Oxides OEM – Original Equipment Manufacturer PM – Particulate Mater PN – Particle Number SCR – Selective Catalytic Reduction VGC – Variable Geometry Compressor VGT – Variable Geometry Turbine VVT – Variable Valve Timing 1.Advanced boosting technology to meet future heavy duty diesel engine requirements A Banks. Changes in engine technology _______________________________________ © The author(s) and/or their employer(s). INTRODUCTION As a result of legislation in the USA and market pressures in all regions. This paper investigates the possibility of existing that the highly efficient technology used in larger turbo machinery may be scaled down to HDD frame sizes. 2014 241 . high pressure ratio. C Such Ricardo UK Ltd. single stage turbocharger is identified. future heavy duty vehicles will need to decrease their Green House Gas (GHG) emissions including CO2 by around 20% over the next 10 years. R Cornwell. 000 km. developments are in hand to introduce measures to control CO2 or fuel consumption. Medium Duty (MD) vehicles have GVW < 15 tonne.000 km per year and have a life before first major overhaul of up to 800. are fitted with engines of about 4 – 8 litre. Within the EU. The expected growth of the European Medium Duty (MD) and Heavy Duty (HD) vehicle market is shown in Figure 1. This paper summarises the expected developments in HDD engine and associated boosting system technology. HD vehicles cover up to about 200.6 Million km. 242 . 2. This may lead to the promulgation of standards on CO2 emissions in a timeframe indicated in Figure 2. The European Commission has sponsored the development of a methodology for calculating CO2 emissions from a wide range of heavy-duty vehicles [3]. trucks and vans carry about 75% of freight transported over land. EU27+EFTA3 Sales and Parc Forecast Commercial Vehicle Parc [Millions] Annual Sales [Thousands] 450 40 400 38 350 36 MD Sales 34 300 32 250 30 HD 200 Sales 28 150 26 100 24 Total Sales 50 22 0 20 2004 2006 2008 2010 2012 2014 2016 2018 2020 Total Parc Year Figure 1: European annual medium duty and heavy duty truck sales including projection until 2020 [2] The US annual sales volumes are similar to Europe. whilst China is producing in excess of 1 Million HDD engines / trucks per annum.make up a substantial part of this reduction and boosting systems have a significant part to play. EURO VI LEGISLATION AND BEYOND Euro VI legislation was introduced for heavy duty vehicles in January 2014. In parallel. Heavy Duty (HD) vehicles are typically defined by a Gross Vehicle Weight (GVW) from 15 tonne up to 40 tonne (certain countries allow 60 tonne) and are fitted with engines from about 8 litre to 16 litre displacement with power levels up to 750 HP (559 kW). which is intended to ensure that the requirements of Euro VI are met on the road. with growth rates exceeding those of Europe and the US. delivering 18 billion tonnes of goods per year [1]. travel up to 150. focus is likely to be on In Service Conformity. In the next period.000 km per annum with a target life before first major overhaul of up to 1. which is in turn defined by the engine’s design Maximum Cylinder Pressure (Pmax) limit Exhaust Gas Recirculation (EGR) rate: the lower the EGR rate which is needed to achieve the required engine-out NOx emissions. which offers potential advantages in terms of CO2 emissions. In recent years. in 243 . another strand in engine development over the next decade will be the utilization of alternative fuels and. EGR rates have tended to reduce as SCR efficiencies have been increased by thermal management of exhaust systems. down to 0. further reduction in NOx will depend on the results of air quality monitoring. then no further tightening of NOx limits can be expected.e. EGR is fitted to the majority of engines. Figure 2: European emission legislation and expected fuel consumption legislation Due to the trade-off between NOx emissions and CO2 emissions. Emissions control technology is common to all Euro VI HDD engines: Selective Catalytic Reduction (SCR) for NOx control and Diesel Particulate Filters (DPF) for control of particulate mass and particle number emissions. the lower the boost pressure requirement Combustion system: this influences the minimum Air/Fuel Ratio which can be used to achieve very low engine-out soot emissions Other factors which have a secondary influence on the turbocharger are the pressure drop across the engine (back pressure of the exhaust aftertreatment) and whether the engine has open or closed crankcase breathing. fuel price and energy security.Although not the subject of this paper. EURO VI ENGINE TECHNOLOGY AND BOOSTING SYSTEMS The factors which most influence an engine’s boosting system requirements are: The highest engine power and torque levels over the speed range. a major question is whether a further round of NOx legislation will be introduced beyond Euro VI. 3. natural gas. in particular. This trend has enabled.02 g/kWh. It is worth noting that the California Air Resources Board has proposed a voluntary NOx limit which represents a 90% reduction below current US NOx limits i. new HDD engines have been designed to a Pmax limit of 200 – 210 bar (actual operation about 10% below the design limit) and this is expected to increase towards 230 – 240 bar before the engine reaches a limit dictated by increasing engine weight and cost. In Europe. If as expected the Euro VI regulations cause a lowering in ambient NOx. lower displacement engines in certain applications in the interests of CO2 reduction. Other applications of mechanical turbo compound systems have been introduced on a US 2010 engine [9] and a Euro VI engine [5]. the EGR cooler has been deleted. the future trends for HDD engines are shown below in Figure 3. Turbocharger technology such as turbo compounding. The Scania DT11 engine originally launched in 1991 achieved a minimum BSFC of 186 g/kWh which equates to a BTE of 45. notably: Single stage turbocharger with wastegate and hot EGR [5] Asymmetric. 244 . and in another case. The long term target is to increase this above 50%. In terms of engine power rating. single stage turbocharger used to drive EGR from one branch of the exhaust manifold [6] Variable Geometry Turbocharger (VGT) with [7] or without cooled EGR [4] Two stage turbocharger with inter and after cooling and cooled EGR [8] All on-highway HDD engines which use EGR employ the high pressure route where EGR is taken from upstream of the turbine. it is likely that specific power will increase from the current level (typically 26 – 32 kW/litre) towards 34 – 36 kW/litre. The alternative. which recovers some of the exhaust energy by means of a turbine downstream of the turbocharger. can help to reduce BSFC on long distance trucks. EGR to be deleted altogether in favour of high efficiency SCR (c. Specific ratings Max cyl pressure Engine rated speed EGR rates BSFC (long term) SCR efficiency Fuel injection pressure Turbocharger pressure ratio >3. 45% (corresponding to about 188 g/kWh diesel). A range of boosting systems is employed on HDD engines.one case.5 Increasing Decreasing Figure 3: Future HDD engine trends 4. This will offer the potential to replace larger engines with more compact. In summary. Around 5% improvement in fuel economy is feasible for long distance trucks. OPTIONS FOR REDUCED CO2 EMISSIONS Currently the highest Brake Thermal Efficiency (BTE) of a Euro IV/V engine at its most efficient operating point is c.5%. to combine the beneficial effects of hot EGR on exhaust temperature at part load and high efficiency SCR at higher loads [5]. low pressure route where EGR is taken from downstream of the DPF is used in some light duty engines but is not considered sufficiently durable on HDD applications. 96% efficient) [4]. in the longer term. The effect of down speeding is to reduce pumping losses and engine friction. allowing higher pressure ratio and reduced flow range compressor maps to be developed. there may be a case for waste heat recovery in applications which consume large quantities of fuel.Another method of reducing vehicle CO2 is down speeding the engine. The introduction of variable valve timing on the inlet valve closing (known as Miller cycle) is being investigated at the research stage. If the Miller principle is used. Alternatively. A recent paper [12] showed that. on-highway truck was about 2 years in the USA and 1. Waste heat may be recovered in the form of mechanical shaft power by turbocompounding.5:1 to maintain the same trapped AFR. the payback time of an ORC system on a heavy duty.5 years in Europe. and also the adoption of new transmissions. Initial results appear promising in terms of potential CO2 benefits [10]. which has proven durability in on-highway applications and has the added benefit of a power increase. this will require an increase in boost pressure ratios above 3. Figure 4 shows an example from an ongoing project [10]. at current fuel prices. the success of waste heat recovery systems in developed markets will depend on payback times. Assuming acceptable reliability and durability. The down speeding is made possible by increasing the BMEP of the engine and achieving constant power at a lower engine speed. Figure 4: Down speeding on a medium duty diesel engine [10] In this case it has been estimated that down speeding by about 18% has the potential to improve fuel consumption by about 2% over a typical driving cycle of regional long distance truck. 245 . in which case the transmission control in the vehicle maintains the engine in the most economical speed range between 900 and 1400 rev/min. the Organic Rankine Cycle (ORC) utilises heat from the EGR cooler and/or the exhaust downstream of the aftertreatment to generate mechanical or electrical energy. Despite its high cost and complexity. The down speeding concept has been introduced on a Heavy Duty vehicle [11]. The reduction in engine speed also reduces the flow range requirements from the turbocharger. the electrical turbocompound approach [13] is a convenient approach.Waste heat recovery can also be converted into electrical power by electrical turbocompounding or thermo-electricity. The high pressure ratio compressor is achieved using a vaned diffuser. one manufacturer produces a range of turbochargers which can achieve >5:1 pressure ratio from a single stage aluminium compressor. incorporating an asymmetric turbine housing. Until recently. TURBOCHARGER INDUSTRY REQUIREMENTS The current solution for high boost pressure ratio requirements (>4. The turbocharger compressor [14] is a single stage unit. the down speeding trend reduces the speed range of the engine which in turn reduces the flow range requirements of the turbocharger. As noted above. 5. and the use of an aluminium compressor wheel is made possible by additional compressor blade cooling. An interesting example of a new in- house manufactured turbocharger is Figure 5: High pressure ratio shown in Figure 6. to maximise turbine efficiency.8:1 pressure ratio. with or without inter-cooling.2) is to use 2 stage turbocharging. 246 . on highway HDD turbocharger development was focused on driving EGR. future engines will have reduced EGR levels or no EGR at all. Compared with the single stage turbocharger. This will allow the turbocharger designer to re-focus on overall turbocharger efficiency. developed for larger non-road applications. The reduction in EGR will allow the complete use of exhaust pulse energy. For generator sets in particular. If we look to the larger frame size machines used on non road applications. As noted above. Certain developing markets such as China are reluctant to adopt solutions such as VGT and two stage which are seen as unacceptably complex and expensive. The challenge for the turbocharger industry is therefore to develop very high pressure ratio single stage turbochargers. Figure 5 shows a very high pressure compressor (140mm diameter) with a peak of 80% compressor efficiency at 5. this has a major impact on the system cost and introduces packaging and complexity issues. which will allow similar compressor design methods to be introduced on HDD engines. 2 4.2:1.4 4.4 4.2 3.1 Turbine Wheel and Housing Design Turbine wheel design will focus on low specific speed wheels. The peak turbine efficiency of 71% is achieved with a radial design wheel and asymmetric turbine housing.4 3. FUTURE TURBOCHARGER TECHNOLOGY 6. These data are for sea level conditions only and assume a 13 litre HDD engine with standard valve timings.Ratio Pmax bar vs Comp P.Ratio 4. the maximum cylinder pressure rises to ~240 bar. Turbine housing design may move away from VGT and back to divided designs (both symmetric and asymmetric) to maximise pulse energy into the turbine.6 3. driven by increased mass flow (associated with lower EGR). at which point the compressor pressure ratio is ~4.4 3. Figure 6: Current compressor and turbine maps for Daimler OM 471 engine [6] The compressor has a very wide flow range and incorporates a ported shroud housing design.1:1.8 3. used to drive EGR efficiently. Sector divided. New low cost variable turbine technology may be introduced also on niche applications. and reduced engine speeds. BMEP bar vs Comp P. As the rating increases up to 30 bar BMEP.6 3. including an allowance for altitude operation to about 2000m without engine de-rate. The rating of 25 bar BMEP is currently the limit for a single stage turbocharger. the peak island efficiency of 80% is achieved at a low pressure ratio (~2.8 3.0:1) and the maximum pressure ratio achieved is 4. 6. as used by larger 247 . Figure 7 shows the future HDD engine trend for BMEP (at peak torque) and Pmax as a function of compressor pressure ratio. turbine housings. rather than meridional divided.2 3 3 24 26 28 30 32 190 200 210 220 230 240 250 Figure 7: BMEP trend at peak torque and P max trend plotted against compressor pressure ratio requirements Current BMEP levels of Euro VI engines at peak torque are about 24 – 26 bar.2 4 4 3. may improve turbine efficiency by better separation of engine pulses. 6.VTG (single stage) 6 -2stage with VTG 1 0 0% 100% 200% 300% Cost Increase Relative to w/g * Estimated using Ricardo analysis Figure 8: Turbocharger technology – compressor pressure ratio vs.2 Compressor Wheel and Housing Design Compressor wheel designs will continue to develop low specific speed wheels. Abraidable coatings can assist in elevating compressor efficiency: this technology has been developed for light duty applications. cost increase (baseline wastegate unit) 248 . To offset this. 6 Opportunity new HP ratio single stage unit 4 6 2 5 Pressure ratio capability 1 3 5 4 3 1 –W/G (single stage) 2 –New turbo HP (single stage) 3 – Low cost Variable turbine (single 2 stage) 4 – 2stage with fixed geom 5. in both aluminium and titanium. and further developments of the port are being developed. based on typical HDD production volumes. Higher pressure ratios will continue to drive developments of the seal design: this will also assist in reducing crankcase pumping work on the engine. alternative EGR concepts may be realised. Another technology which will continue to assist in increasing the turbocharger overall efficiency is low friction bearings (including ball bearings) which are already introduced on a medium duty application [17]. Ricardo’s investigations into a dual circuit EGR system [15] where EGR was taken from both high and low pressure sources showed that good fuel consumption could be realised whilst maintaining the ability to drive high levels of EGR. COST OF TECHNOLOGY Figure 8 shows the estimated trade-off of alternative boosting systems in terms of pressure ratio capability and the cost compared with a wastegated single stage turbocharger. Wheel-to-housing clearance for compressors will become more and more important as compressor pressure ratio increases and wheel blade height reduces.turbochargers. Ported shrouds will continue to be used. such as low pressure EGR. 7. Variable Geometry Compressors (VGC) have been developed and offer improvements in map width: this technology may also be applied to the vaned diffuser [16]. Increasing the overall turbocharger efficiency will reduce the ability to drive EGR. and are needed if we are to deliver higher specific engine ratings and >50% engine Brake Thermal Efficiency (BTE). and an opportunity exists for a high pressure ratio single stage unit. LMC Automotive. et al “Common rail Integration on PACCAR Heavy Duty Truck Engine” Aachen Colloquium. et al “The 2013 Daimler DD15 Engine – Embracing NAFTA’s 2014 GHG and 2013 OBD Regulations with Benchmark Fuel Economy” Aachen colloquium.. September 2012 5. Technologies used on larger turbochargers are likely to be adopted for smaller frame sizes. CONCLUSIONS New HDD engines will use very high BMEP >30 bar. Sep 2007 9. Gothenburg. Germany. and introduce high specific ratings up to about 36 kW/litre. providing significant benefits in terms of cost and package. REFERENCES 1. Duesseldorf. “Progress in CO2 and Fuel Efficiency Goals and Measurement Methodologies” presented at 9th Diesel Emissions Conference & AdBlue Forum Europe 2013. 2 stage solutions are needed with associated cost and packaging issues. Ricardo Analysis. above this. et al “Turbocharger Concept for Current and Future Commercial Vehicles” 12th TU Dresden Technical Conference on Supercharging. Technology level 2 in Figure 8 (high pressure ratio single stage turbocharger) offers a high pressure ratio capability at a relatively low increase in cost when compared to the baseline. October 2013 7. Significant single stage turbocharger efficiency gains are possible. two stage solutions will be required. de Kok. April 2008 249 . single stage turbochargers are an attractive alternative to 2 stage turbochargers. October 2013 8. Gruden. This requirement is especially necessary if we consider turbo compounding solutions.The current single stage turbochargers are limited to 3. Heil. 8. B. von Hoerner. ACEA.5:1 pressure ratio at sea level. June 2013 4. Automotive World 3. S. M. Signer. Such a turbocharger would offer an interesting potential solution to the problem of the engine manufacturer who wants to extend the power range of his engine whilst maintaining commonality of the boosting system. et al “New Daimler Heavy Duty Commercial Vehicle Engine Series” 29th Vienna Engine Symposium. Hausberger. European Automobile Manufacturer’s Association ACEA publication “The Automobile Industry Pocket Guide 2013” 2.. R. The new ratings will incorporate high P max capability up to 230-240 bar. Nakano. The new generations of turbocharger will be combined with other technologies such as variable valve timing to achieve high specific power and improved CO2 emissions. Lyon. The trend for boosting technology is to increase the overall pressure ratio requirements > 4:1 in combination with very high overall efficiency. D. “Volvo Group Engine Evolution to Meet Euro VI Emission Standard” SIA Conference. D. These ratings will force further efficiency gains on the turbocharger. P. 9. November 2012 6. “Meeting Euro VI and EPA10 Legislation without EGR” SAE HDD Emissions Conference. High pressure ratio. ABB.. J. A. Such CH. P.com/our-technologies/ball-bearing/ .www.J. S. presented at Institution of Mechanical Engineers. D.use on Hino medium duty truck 250 . Tett. “The Role of Waste Heat Recovery in Meeting Phase 2 US EPA Greenhouse Gas Regulations” 9th Diesel Emissions Conference. R... Single stage high pressure ratio A100 compressor – www. N Watson and MS Janota Turbocharging the internal combustion engine – variable inlet guide vanes pg 135-136 17. Patterson. M. Niven. McGuire... Edwards. A.com 15.volvo. “Potential reduction of CO2 emissions from Heavy Duty Diesel engines for long distance transport” ATZ Live Conference. Banks. SAE paper 2009-01-1604 14.honeywell. http://turbo.10. “Boosting technology for Euro VI and Tier 4 Final Heavy Duty Diesel Engines without NOx Aftertreatment”. Ludwigsburg 5 & 6 November 2013 11. Andersson. Duesseldorf. “Exhaust Heat Recovery using Electro-TurboGenerators”. May 2010 16.com 12 Stanton.P. June 2013 13. London. Volvo iShift transmission .T.C. North American Technical Center. . this specification no longer reflects the state of the art. First issued in 1989. key issues are identified which may produce misleading findings from turbocharger gas stand performance testing. T-S Turbine combined efficiency Turbine Power Π Compressor Pressure Ratio = Bearing Housing Parameters / Oil Inlet Temp Compressor efficiency = Heat Flux at bearing housing ( )/ diffuser surface Compressor Power _______________________________________ © The author(s) and/or their employer(s). realistic results while also enabling a fair comparison of product performance between manufacturers of turbochargers. USA ABSTRACT The SAE “Turbocharger Gas Stand Test Code” (SAE J1826) has historically defined the best practices for testing of turbochargers on gas stands. ensuring accurate. ex Specific Heat of Exhaust Gas Compressor Reference Pressure Ratio of Specific Heats for Compressor Reference Exhaust Temperature Turbine physical mass flow Cp Specific Heat of Air Φ Turbine Flow Parameter = Ratio of Specific Heats for Air Compressor physical mass flow Π T-S Expansion Ratio = Compressor Corrected Flow = . additional practices beyond the scope of J1826 must be followed. .Considerations for gas stand measurement of turbocharger performance J B Schwarz. D N Andrews BorgWarner Turbo Systems . / . To obtain consistent thermodynamic mapping of turbochargers. NOMENCLATURE Compressor Parameters Turbine Parameters Compressor Inlet Pressure Turbine Inlet Pressure Compressor Inlet Temperature Turbine Inlet Temp Compressor Discharge Pressure Turbine Discharge Pressure Compressor Discharge Turbine Outlet Temp Temperature Cp. Using empirical gas stand test results as a basis. 2014 253 . the surge line represents the minimum operable flow rate of the compressor over a range of speeds. repeatable. Aspects of gas stand configuration and test operation are not specified in sufficient detail in the specification. The accuracy of performance figures is sensitive to the design of the gas stand itself as well as the selection and location of measurement devices. (1) The field of turbocharging has evolved considerably since this specification was outlined in 1989. Practically. (2) Additional reading is provided in Mai’s study of significant parameters in gas stand turbocharger testing. The surge event is triggered by stalled flow conditions within the compressor stage. Turbine and compressor flow measurement devices are external to the components shown in Figure 1. The volume and geometry of the compressor inlet and discharge ducting play a significant role in 254 .1. Stage performance and map width can be altered through manipulation of these parameters. (3) Thermodynamic mapping depends on accurate measurement of temperature and pressure changes across each stage as well as measurement of flow through the stage. 1.1 Overview of Gas Stand Hardware The gas stand provides a steady-state hot gas test bed on which measurements of turbocharger stage performance can be made. however. INTRODUCTION The SAE “Turbocharger Gas Stand Test Code” (SAE J1826) was created as a guideline to ensure consistent thermodynamic mapping of turbochargers across the industry. A more thorough discussion of modern gas stand design and capabilities is provided by Young and Penz. as a system instability. and accurate turbocharger performance test results. surge is sensitive to numerous aspects beyond the compressor itself. and shortcomings in SAE J1826 undermine the original purpose of ensuring consistent. Figure 1: Typical measurement sections and adapters found on a gas stand used for turbocharger performance measurement 2. SURGE LINE AND MAP WIDTH Surge is a compressor instability phenomenon marked by bulk flow reversal across the stage. Figure 1 shows a typical layout of adapters and measurement sections required for temperature and pressure measurement on the gas stand. 08 0. Figure 2: Surge line shift due to change in system inlet restriction 3.0 0. changes in result in an inverse effect on the turbine side.20 0.06 0. . (6) Given the dependence of surge on duct geometry.8 3. HEAT TRANSFER EFFECTS Turbocharger stage efficiency calculations depend on accurate calculation of enthalpy change across compressor and turbine stages. This means that different turbochargers can experience surge at the same compressor flow conditions for different underlying reasons. a more representative surge line can be obtained by using application ducting. compressor inlet restriction due to ducting has been observed to have a marked effect on the surge line. various turbochargers may respond differently (in terms of surge) to changes in the system. the only reliable method for obtaining a consistent surge line is to maintain a consistent compressor ducting setup across all turbocharger testing.18 0.04 0. as seen in Figure 2. This provides a repeatable surge line and generally allows for a quantitative comparison between turbochargers. (4) SAE J1826 makes no provisions regarding the design of compressor ducting.26 . is related through a power balance.2 1.10 0. 3.14 0. Heat flux into or out of a compressor stage will cause a change in the apparent efficiency of the stage. However. a range of non-adiabatic effects can occur.16 0.22 0.00 0. 255 . . the onset of surge is particular to whether the stall is rotating or stationary and whether it occurs at the impeller or in the diffuser.0 2.12 0. As a result.determining whether the system will be destabilized. causing a surge.24 0. as seen in Figure 3. As surge is triggered by a stall event.4 1.6 2.4 Surge line with compressor inlet restriction Pressure Ratio (Total-to-Static) 3.8 1. This result was obtained by testing a single turbocharger with multiple levels of inlet restriction. Although the system is assumed to be adiabatic. Specifically. Similarly.02 0. (5) Pre-swirl or counter-swirl inlet vanes can also be employed to shift the operating space of the compressor map. therefore. it has been observed that a suction-side butterfly throttle valve can move the surge line. a function of the pressure and temperature rise across the stage. Heat transfer between bearing housing & lubricating oil T1 influence T3 influence Conductive heat Radiative. 256 . A number of phenomena can change the thermodynamic balance. some dramatically affecting stage performance results. a robust technique is to use the turbocharger power balance to obtain the combined aerodynamic/mechanical turbine efficiency. The turbine power equation takes a similar form. non-adiabatic heat transfer dominates relative to adiabatic compression and expansion. Here. The compressor power is described by / Π 1 The power required by the compressor is taken to be equal to the power delivered by the turbine. conductive flux through heat flux from turbine bearing housing housing Convection at diffuser face Figure 3: Examples of turbocharger heat transfer phenomena 3. SAE J1826 does not provide specific guidance for constraining parameters related to heat transfer.Generally.1 Compressor-Turbine Power Balance Compressor efficiency can be calculated by determining the stage enthalpy change. To find the turbine efficiency. these thermodynamic effects are found to have the greatest influence at low flow rates and for small turbochargers. Insulation was then progressively removed. 1 − (1/Π ) . .06 0. the efficiency near surge increased by up to 4%.8 -1 1. as compared with insulated case (left).1 0.16 0. is inversely dependent on . .04 0. This dependence on combined with the additional turbine side parameters ( . SAE J1826 specifies that the turbocharger should be insulated for testing. When tested without insulation.2 Insulation Of the many phenomena that can affect heat transfer on a gas stand. .6 -2 1. .18 Corrected Compressor Mass Flow (kg/s) Turbo with insulation Figure 4: Percent change in T-T compressor efficiency when run without insulation. the equation can be solved for . thereby greatly reducing the influence of ventilation systems on test results.14 0.12 0.08 0.2 1 2 0 Turbo without insulation 1.4 -3 -4 0. In this representation.) forces the . many test facilities opt to test turbochargers without insulation. uncertainty to always be higher than the uncertainty. but there was a pronounced effect on smaller turbochargers. . 3. however. etc. Refractory insulation with a reflective backing can significantly improve the accuracy and repeatability of turbocharger stage efficiency measurements. Figure 4 shows Total-to-Total (T-T) results for a small turbocharger typical of passenger car and light-duty diesel applications. includes both turbine aerodynamic efficiency and bearing system frictional losses. it can be seen that . 4 2. Assuming = . = . and compressor cover.6 3 2. Baseline performance was determined for a turbocharger with full insulation on the turbine inlet. test turbocharger shown without insulation (upper right) and with insulation (lower right) 257 . while choke efficiency decreased.4 2 Compressor PR (T-T) 2. turbine housing. Insulation was found to have a reduced impact on medium to large turbochargers typical of commercial diesel applications. insulation has perhaps the greatest influence. Insulation prevents radiative heat transfer from hot surfaces and reduces convective transfer of heat between turbocharger components and the test cell. Experiments were conducted to quantify the impact of insulation on several turbocharger frame sizes. 75 100% Cooling Water Flow 0.08 0. The authors’ recommend that a test facility tightly control test cell air temperature and not use any insulation. In order to project realistic estimates of stage performance. however the type and method of insulation has been observed to cause and . particularly for small frame passenger car turbochargers.24 0. For aerodynamic development. the authors recommend that no cooling water be supplied to the turbocharger.16 0.65 0. both insulated and un-insulated test methods have their drawbacks. Since . Accurate control of test cell temperature will require a dedicated cell HVAC system. this increase in results in a decrease in . There does not appear to be a clear consensus within the industry regarding insulation practices.04 0. is computed via a power balance.70 0.12 0.50 0. 0. From a repeatability standpoint.28 (kg/s) .55 0% Cooling Water Flow 0. variation. This approach eliminates variation caused by differences in operator methods of insulating the turbocharger. Figure 5 shows that for a compressor with a water-cooled bearing housing.20 0. Figure 5: Influence of cooling water flow on compressor efficiency . turbochargers which feature water-cooled components should be tested with cooling water flow rates and water temperatures representative of the application.60 0. . Insulated testing shows little dependence on test cell temperature. the application cooling water parameters can be replicated on the gas stand to provide a closer match with application performance.Of key importance. For application-specific development. Such results are more indicative of raw aerodynamic performance and can be compared fairly against a library of performance results for non-cooled turbochargers. climbed from 58% to 70% with increased cooling water flow. 3. test results from insulated turbochargers cannot be compared directly with un-insulated tests.00 0.80 Compressor Efficiency (Total-to-Total) 0. Un-insulated testing must be carried out with a tight control of test cell temperature to reduce performance variation caused by convective heat loss. 258 . This effect diminished with increasing compressor mass flow rate.3 Cooling Water Water cooling can drive a significant change in apparent stage efficiency by manipulating the enthalpy change across the stage. and the inlet oil temp.5 Turbine Inlet Temperature Varying causes significant changes in calculated values of . the diffuser surface will be significantly hotter than the air leaving the compressor exducer. .4 Compressor Inlet Temperature Measured compressor efficiency can be affected by changes in the compressor inlet temperature. compressor cross section showing convective heat transfer through diffuser wall (right) 3. Figure 6: Influence of compressor inlet temperature and speed on compressor T-S efficiency (left). . and Π . . large frame units are insensitive to changes). RED: T1 = 50 C BLUE: T1 = 35 C BLACK: T1 = 25 Q 370 m/sec 230 m/sec Passenger car turbo with 46mm exducer compressor . The bearing housing diffuser surface temperature is a function of . Cold oil causes an apparent increase in a decrease in . a consistent oil temperature should be maintained throughout all thermodynamic testing. Additionally. especially if the test facility will be running small frame turbochargers. this effect dominates at lower air flow rates and impacts small frame turbochargers more so than larger units. Consistent with other heat transfer effects. a small decrease in .Cold lubricating oil produces an effect on the compressor stage similar to that seen for cooling water. The authors’ recommendation is to tightly control T1 via a dedicated HVAC system. will be seen due to increased bearing system drag due from higher oil viscosity. due to the compressor-turbine power balance. Testing on larger commercial diesel size turbos has not shown this effect (i. 3. The effect of on can be mitigated through control of the inlet air temperature to the compressor. Φ. This convective heat transfer can be a significant percentage of the turbo’s overall shaft power for small frame turbos at low tip speeds. This heat transfer effect occurs mainly in the parallel wall diffuser section and is driven by convective heat transfer from the bearing housing to the airflow.e. At low tip speeds. as seen in Figure 6. Therefore. This affects the convective heat transfer into the airflow in the compressor diffuser section: 259 . Variation in changes the heat flux into the bearing housing. As is raised. In addition. . TURBINE TOTAL TO STATIC PERFORMANCE AND TURBINE DISCHARGE MEASUREMENT STACK SIZING It is standard practice for automotive turbocharger manufacturers to report turbine performance on a Total-to-Static (T-S) basis. rises. Small frame passenger car turbochargers will show the most efficiency variation when is changed. This creates an increase in . 260 . It is challenging to accurately determine . the Π range for a given speedline will increase. will rise due to the higher calculation of compressor power. If a power balance is used to find . A large stack diameter lowers turbine outlet velocities and creates more static pressure recovery. .45. a reduction in Π .3 to 1. . and . The Π shift causes the Φ values to slide up and to the right along the Φ curve. the allowable range for measurement stack diameter / turbine wheel inducer diameter could be 1. To minimize these various effects. on a T-S basis stems from its sensitivity to the turbine discharge measurement section diameter. This requires the total temperature and pressure properties at the turbine inlet and only the static pressure at the turbine outlet. These changes are most significant at low compressor speeds and flow rates. can be greatly manipulated by changing the discharge stack diameter.6 A Final Note Regarding Compressor Heat Transfer Effects When discussing heat transfer effects and their impact on . it is recommended that a consistent value be used for all thermodynamic testing. shifting to the right on the turbine map. As is increased for a given . Total measurement techniques (such as pitot probes and total temperature probes) are not suitable for high-volume automotive turbocharger testing. For a given turbo. and an increase in the calculated . For application-specific testing. For aerodynamic development. SAE J1826 does not provide guidance regarding sizing of turbine outlet measurement stacks. therefore the velocity and density at the turbine discharge will not be known. it must be noted that none of these cause variation in the measured Π . These changes only impact and the calculated – not Π . a measurement stack should be selected so that a consistent ratio is maintained between the stack diameter to turbine wheel inducer diameter. 3. When running turbine performance testing to a set of target turbine corrected speeds. and the resultant calculated decreases. the compressor operates at higher compressor tip speeds and shaft power as is increased while operating at a certain set of turbine corrected speeds. the reported . The danger of reporting . a customer-specific outlet pipe diameter can be used. For example. 4. . changing will cause significant changes in the turbine operating points. at the turbine discharge because of the variable rotational component of the flow velocity and the large temperature gradient that exists in the flow. Figure 7 shows performance results with two different sizes of turbine discharge stacks for a medium size commercial diesel turbo. the calculated value of . 6 5.4 1. when varying . as seen in Figure 8.2 2. The effect of varying on . control over is easily accomplished via an exhaust throttle valve placed downstream of the turbine discharge measurement section. 261 . two stage architecture. and Φ . At lower flow rates. and Φ.6 3. TURBINE DISCHARGE PRESSURE Turbine efficiency can often be influenced by . In this scenario. . the change in caused the turbine operating point to shift.0 1. the turbine showed elevated .4 3. Application is usually close to atmospheric pressure but can be elevated in specific cases due to exhaust system piping. Testing for aerodynamic development necessitates that turbine performance be obtained at consistent conditions to permit direct comparison with an existing library of turbine performance data. Testing on smaller passenger car size turbochargers has shown little to no change in .4 Expansion Ratio P3/P4 (Total-to-Static) Figure 7: Influence of turbine discharge measurement stack diameter on turbine efficiency 5.575 0.550 0.525 1.2 4.650 68mm stack Turbine Efficiency (Total-to-Static) 0. following the general shape of the speedline curves for . Most gas stand testing is done at a nominal that is not representative of the actual application.8 4.600 0. and after-treatment. as was increased. For application-specific testing. at higher Π .0 3.0 5. Testing conducted on a medium-sized commercial diesel turbocharger showed that aggressive changes in turbine outlet pressure produced a change in .625 0. Therefore.675 89mm stack 0. application should be considered when interpreting . Ideally the valve would be adjusted to approximate the levels expected in the application. the change in produced a marked shift in . is not universal for all turbine wheel designs.8 2. However. results. 0. 57 tip speed High P4 0.61 0.59 224 m/sec Normal P4 Moderate P4 0. the bearing system losses are incorporated into the reported . but the effect is somewhat more for very small turbochargers. In addition.5 3 3.65 360 m/sec Turbine Efficiency tip speed 456 m/sec 0.67 0. .5 4 4. In this circumstance. 0. Decreasing Oil Viscosity Turbine Efficiency Passenger car turbo with 29mm exducer compressor Expansion Ratio Figure 9: Reducing oil viscosity causes an apparent increase in turbine efficiency for a small turbocharger 262 .5 Turbine Expansion Ratio (T-S) Figure 8: Effect of back pressure on turbine efficiency 5.69 0.55 1 1.5 2 2. These losses are generally less than 1%. .1 Bearing System Losses Due to Oil Temperature and Viscosity When computing turbine combined efficiency using a compressor-turbine power balance. oil temperature can affect the system through two mechanisms: an increase in temperature causes a reduction in viscosity and also changes the bearing housing heat flux.63 tip speed 0. oil viscosity has a modest effect on . For testing of small frame turbochargers. higher-flowing 263 . Environmental heat transfer effects are most significant at low flow rates and on smaller turbochargers. the test facility would use no insulation but would employ a dedicated HVAC system to control test cell air temperature. Stage performance obtained using water cooling cannot be compared directly with existing non water-cooled performance data. is sensitive to the diameter of the turbine measurement stack relative to the wheel size. Frequently. hardware configuration. Future gas stand development promises greater accuracy for engine simulation. A gas stand turbocharger testing facility generally serves two unique customer types: aerodynamic development groups and application-specific customers. higher can cause an apparent increase in . but oil viscosity influence was found to be insignificant aside from small turbochargers typical of passenger car applications. increases as a function of the stack diameter. . . Several environmental factors impact the thermodynamic balance of the turbocharger. Results must be taken with a degree of uncertainty. 6. If the oil type is changed. The compressor map shape and surge line are affected by the design and material of the compressor ductwork. can be shifted by approximately ±1% for a very small turbocharger as a function of the viscosity and formulation of the oil. especially for large. Although on-engine testing currently provides the most realistic view of turbocharger-engine system performance. these caveats also apply. Insulation has perhaps the most effect on apparent compressor and turbine stage performance. especially the resistance of the inlet ducting. . A test facility must maintain a consistent practice to permit direct comparison with a data library. . and measurement techniques. CONCLUDING REMARKS This paper has showcased several significant phenomena which can influence turbocharger test results. when comparing product performance between turbocharger manufacturers. This effect is most pronounced at the lower speedlines and rapidly diminishes with increasing turbine flow. Water cooling should not be supplied to bearing and turbine housings during testing for aerodynamic development.As seen in Figure 9. The needs of these customers will often dictate certain testing conditions and practices. Similarly.” there are many caveats not provided for in this specification. unless the testing facility can provide specifics of the test procedure. oil inlet temperature and T should be kept at constant values. the accuracy of gas stand mapping has improved significantly due to knowledge of the factors discussed in this paper. To permit comparison of stage performance. When thermodynamic test results are obtained in accordance with the SAE “Turbocharger Gas Stand Test Code. it is recommended that a consistent oil viscosity and formulation be maintained. Other frame sizes of turbochargers were tested. gas stand performance maps are used to produce engine power simulations. Additionally. the authors are aware of several facilities which test turbochargers without insulation. Although SAE J1826 calls for the turbocharger to be insulated during testing. thermodynamic performance should be compared against reference data to evaluate the impact of the change on stage performance. and other effects exist beyond the scope of this paper. To ensure a consistent map shape. the compressor duct configuration must be kept as similar as possible between multiple gas stands. Ideally. D. Aerodynamic development work is best suited by maintaining a consistent ratio of turbine stack diameter to wheel diameter for all turbine performance testing. the turbine stack diameter should closely match the application’s downpipe diameter. Texas. Eng. pp. SAE J 1826. 3. Turbocharger Gas Stand Test Code. 6. M. Berlin : VFI / Technische Universität Berlin.. Mai. Stall. E. : ASME J. Boyce. 4. 2. SAE Engine Power Test Code Committee. The Design of a New Turbocharger Test Facility. and Surge. Parameterstudie zur Turbolader-Kennfeldvermessung. Small passenger car turbochargers show a sensitivity to oil viscosity and temperature.stages. Stage Stall. SAE International Congress and Exhibition.P. Japikse. pp. Surge and Rotating Stall in Axial Flow Compressors.M. 5. pp. 98. Vol. pp. for Power. et al. 01. April 1976. 1990. Holger. 1-13. Texas. Greitzer. s. REFERENCE LIST 1. Proceedings of the Twelfth Turbomachinery Symposium. . 190-198. Michigan. For application-specific testing. Where . is found via a power balance approach. Practical Aspects of Centrifugal Compressor Surge and Surge Control. factors which affect the bearing system frictional losses. Detroit.A. M. 1981. 1983. and D. 264 .l. USA : SAE International. March 1995. 2009. College Station. USA : Texas A&M University. 40-55. USA : Texas A&M University. College Station. Penz. Young. 147-173. these factors will affect calculated .Y. Part I: Theoretical Compression System Model. Proceedings of the Tenth Turbomachinery Symposium. France c Hochschule Weingarten. Limitations and possibilities of testing repeatability are explained and the critical factors in the quality of turbocharger testing are discussed. Germany ABSTRACT This article is focused on the quality criteria of the measurement results from Turbo Charger Test Benches. The testing facility can be regarded as a complex measurement system. A Rinaldi b.Critical aspects in turbocharger testing H Bolz a. Total-to-total compressor isentropic efficiency Π . References are made to the method of “Measurement System Analysis” as far as it is compatible to the complexity of turbo charger measurement. The results are affected by several parameters and boundary conditions. Germany b CRITT M2A. . Compressor corrected rotational speed η . Some of them can be influenced by the test system itself. The aim is to define clearly and precisely the quality criteria for the repeatability and reproducibility of the measurement results of turbo charger test benches. others are affected by the environmental conditions. In particular this paper discusses the conditions which can be influenced by the test bench (Measurement and Control) and defines a test procedure to verify the quality of the test results. NOMENCLATURE Symbol Physical value . in order to give reliable measurement results for later use in engine process simulation and engine control unit calibration. Total-to-total compressor pressure ratio Volumetric-flow Mass-flow p Pressure Q Heat flow T Temperature φ0 Relative ambient humidity _______________________________________ © The author(s) and/or their employer(s). 2014 265 . An analysis of these phenomena is presented. Intensive study of repeatability and reproducibility is reported and some considerations of linearity and long term stability are presented. This work gives an analysis of these aspects to support testing procedures as they are known to be important in turbocharger development and comparison. A Kaufmann c a Kratzer Automation AG. Compressor corrected mass-flow u . Following the increasing demand for turbochargers and the speed up in research activities. The temperature is obtained by heat supply from a 200 kW natural gas burner.1 Basic Boundary conditions All the experimental results presented in following pages have been obtained by two different test benches both produced by Kratzer Automation AG. Some recommended procedures can be found in the Turbocharger test stand test code (1) and related literature (2) (3). Three temperature probes type Pt100 are inserted in the fluid vane both at compressor inlet and outlet. These topics are well known issues in the turbocharger environment. 266 . A lubrication rig is used to ensure the correct operation of the bearings. it is intended to make public some studies with the specific goal of improving the knowledge in this field. Figure 1 shows the principal points where the measurements used for the thermodynamic calculations and gives a rough overview of the heat flow in the test object.1 INTRODUCTION The characteristic maps of turbochargers [TC]. The turbocharger turbine is fed with compressed air at defined temperature while the compressor can operate at a controlled pressure level. Air humidity. Care is taken to compare results from exactly the same setup. The flow rate measurement can vary from 0.25 kg/s. This justifies the importance of measurement and the corresponding boundary conditions. engine control unit calibration. Results are obtained following a specific procedure of using MSA (Measurement System Analysis). The overall study focuses on some critical aspects of turbocharger measurement and works out basic quality criteria on measurement results. Three thermocouples type K are inserted at the turbine interfaces. are the main sources of information to quantify the performance of the machine. research activities and several other tasks in internal combustion engine development processes. Recently E. Standard pressure sensors are placed at the interfaces on a stabilization ring which physically averages three pressure tappings on the wall of the flow. It is characterised by the availability of two independent supply lines for compressor and turbine. Guillou (8) reported on the uncertainty and measurement sensitivity of turbocharger compressor gas stands. for engine process simulation. Guillou. The most important entities are known and still little literature is available. adaptation parts and measurement pipes are used in the study. One single set of turbocharger. This paper will discuss the impact of different parameters starting from the analysis of the different parameters affecting the measurements. They constitute the basic input for application of a turbocharger. some work on measurement uncertainties has been reported by Brun and Kurz (4) and others (5) (6) Chapman and Schultz (7) which provide information on gas turbine field test measurement uncertainties.01 kg/s to 0. Following the approach of E. The turbine hot gas temperature can be regulated from 150°C to 1200°C. Compressor back pressure is controlled by an appropriate back pressure regulating system. test cell air conditions and turbocharger rim speed are acquired with commercial sensors. both for compressor and turbine. A mathematic approach is discussed for comparison of two different test benches with identical equipment with respect to comparability and repeatability. 2 INFLUENCE PARAMETERS AND TEST SETUP 2. The test rig is a continuous flow apparatus particularly suitable for performing investigations on exhaust turbochargers. correlated with the conditions in pressure loss between ambient and the lab. setup of venting reconstruction for the test cell system. 3 Influence caused by production These influence parameters can tolerances and runtime of the specimen be excluded by using a dedicated specimen. such as pressure.2 Classification influence parameters The influence of factors. temperature mostly combined with around the specimen. back pressure of exhaust gas line. 2 Influences caused by the piping of the These parameters generally test system in the test cell such as construction factors and are volume between TC and flow controller. 267 . 4 Measurement uncertainty in every Parameters mainly fixed and sense. affecting a turbocharger map. compressor input. T0. 5 These parameters are generally given Can be influenced by the by the strategy of controlling and correct control and data storage storing results and have to be strategy accounted for a repeatability study. etc. Figure 1: Turbocharger measurement set up 2. can be grouped in at least 5 different classes: 1 Environment of infrastructure 2 Environment of the test system 3 Specimen 4 Accuracy of the instrumentation 5 Test system control Table 1: Explanation of each class of influence factors Class Description Qualification 1 Influences caused by the ambient These parameters are generally conditions of the complete test cell and given by the initial set up of the the infrastructure for conditioning test cell. temperature are determined by the and flow rate uncertainty instrumentation. Modifications are systems such as p0. etc. When comparing results from different facilities, all these effects need to be taken into account. Concerning classes 1 and 2, few things can be done. Different setups will bring differences in the results. As pointed out in Table 1, only some significant modifications allow big improvements. These operations are generally not possible in the running phase of a turbocharger development project. Concerning class 4 a trade-off between sensor capabilities and cost gives the best solution. Regarding classes 3 and 5 every effort should be made to define a standard setup between different labs. With this general statement in mind, the strategy for the repeatability and comparability study is defined with the aim of quantifying the influencing parameters of classes 1,2 and 4. 2.2.1 Test environment In order to rule out the influencing parameters of the environment conditions corresponding to class 1 and 2 as much as possible, the following requirements have been defined: Strong requirement Level of intake temperature at compressor inlet Range of 20..25 °C Level of temperature variation over the test 0,5°C/h < 1,5°C during test Weak requirement Level of cell temperature around specimen Range of 30..40 °C Level of back pressure exhaust gas system 20 mbar (max.) Airflow control based on heat input in the test cell, to avoid extreme airflow around TC 2.2.2 Specimen setup In order to rule out the influence parameters of specimen corresponding to influence parameter class 3 as far as possible, the following setup was adopted: One single physical Turbo Charger KP35 from Borg Warner 4 x Adaptation parts between TC and measurement pipes 4 x Measurement pipes with a diameter of 50 mm Fixed installed rim speed measurement Fixed installed temperature sensors (Pt100 on compressor side/ Thermocouples type K on turbine side) The test TC is prevented from entering surge or choke in order to prevent high wear or damage. 2.3 Test setup All results presented in the pages below are obtained by the use of the same test sequence. The specimen is run on three consecutive measurement points on three speed lines for ten times. The turbine working point is maintained at T3=600 °C inlet temperature. Oil inlet conditions (temperature and pressure) are kept constant during all the tests. The following Table 2 summarizes the compressor working points. Table 2: Measurement operation points Cor. speed [m/s] 230 370 490 0,025 0,035 0,045 Cor. flow [kg/s] 0,050 0,070 0,085 0,080 0,095 0,108 268 3 SENSOR ERROR AND DATA EVALUATION Experimental determination of the thermodynamic properties from measurement is not precise. There are many sources of error as classified below. This makes the measured value different from the real physical conditions. For the development and comparison of turbochargers it is hence crucial to consider the possible errors since they are fundamental to design evaluation. 3.1 Sensor Errors A sensitivity analysis was run on some calculated variables in order to characterise the influence of measurement uncertainties on the calculation of total-to-static pressure ratio and total-to-static efficiency. As the total error of the measured value largely depends on the installation, only the possible sensor errors given in Table 3 are considered since the focus lies on the error propagation. In addition to ambient temperature T0, ambient pressure p0, relative ambient humidity φ0, inlet temperature T1, exit temperature T2, inlet pressure p1, exit pressure p2, the recovery factor for temperature measurement rf and the change in wheel speed uc on one target speed line are considered. Table 3: Considered sensor errors δT0 δp0 δφ0 δT1 δT2 δp1 δp2 δrf δuC [K] [Pa] - [K] [K] [Pa] [Pa] - [m/s] 0.15 40 0.05 0.15 0.15 190 400 0.1 0.1%u The error is computed using a Gaussian type error propagation analysis. A numerical approach (central differences) is used to compute the error due to the individual measurement error. The error computation for the efficiency is given below. ∆ = ∑ ∆ ≈ ∑ Δ ( ) (1) …, ,… …, ,… ≈ Δ = …, + , … − …, − , … (2) For the evaluation of the individual η … , x ± δx , … full data evaluation using enthalpy and entropy correlations of the gas standards of humid air composition was performed. 3.2 Compressor efficiency and pressure ratio maps This error analysis has been applied to the previously discussed nine test points in the compressor map. The computed error is quantified by an error bar in the y direction. The error bar in the x direction corresponds to the 1% uncertainty of the mass flow rate measurement. 269 Figure 2: Compressor map with error bars Figure 3: Compressor efficiency map with error bars In the case of the pressure map an additional error is present when two maps are compared. Since the compressor wheel speed varies slightly along the iso-speed line (δuC), the representation is not completely exact. As the pressure ratio increases as the square of the wheel speed, this error becomes significant at higher compressor wheel speeds and pressure ratios. This error can be quantified by a simple analysis. /( ) /( ) = = 1+ (3) /( ) = 1+ (4) For the circled point on the efficiency and pressure map a sensitivity analysis is carried out. The results of the sensitivity analysis are summarized in Figure 4 and 5. 270 3.3 Sensitivity to individual sensors Figure 4: Compressor pressure ratio sensitivity to measurement Figure 5: Compressor efficiency sensitivity to measurement The largest error due to sensor error in the efficiency map is given by pressure measurement, followed by the temperature measurements before and after the compressor. The measurement of ambient pressure, temperature, humidity, mass flow rate and the recovery factor are of minor importance. The error in efficiency is larger for the smaller iso-speed lines as the relative error in pressure and temperature measurement are much larger. Investment in precise temperature and pressure measurement would hence give the best benefit. A significant error in the pressure map is due to the 0.1% compressor wheel speed error. Longer holding time can cut down this error significantly at the price of longer measurement times. When this error is neglected, the source of error is dominated by the pressure measurement before and after the compressor. The error due to data evaluation with dry air and constant heat capacities compared to full thermodynamic evaluation of enthalpies and entropies contributes to roughly 0.5 % of the efficiency. 271 4 COMPARABILITY In order to quantify the reproducibility error of the same test specimen the same unit will be measured under the same conditions by two different test stands. The comparability error is then defined as the differences between the results. 4.1 Comparability mathematic model The typical map at a turbocharger test stand provides information regarding the functional relationship = , (5) between the signals , and . For the compressor test setup the signals are : corrected Mass flow : corrected rim speed : pressure ratio (total to total) Corrected mass flow and corrected rim speed represent the independent variables, and pressure ratio is the dependent variable that will take values according to the functional relationship , . At the test stand, the independent variables are set by controllers. This is the difference to the sensitivity analysis via data evaluation in section 3 and will be an additional source of discrepancies. Note that the function , mainly depends on the test specimen and is a priori not known. The measured data is gathered at discrete support points. Therefore, data samples , and are considered instead of one data point. In order to provide information on how far the measurement data of one test stand can deviate from another test stand due to the inevitable uncertainties given by the limited precision of measurement equipment and control quality the following considerations have been taken into account. 4.1.1 Measurement Uncertainties The values , and contain measurement uncertainties [mu] ∆ , ∆ and ∆ due to the utilized sensors. Therefore the deviation can be assumed to be limited by uncertainties. μ , − ≤∆ (6) μ , − ≤∆ (7) μ , − ≤∆ (8) Here μ , , μ , and μ , , are the real (unknown) signal values. In the given case, the measurement uncertainties are obtained by means of error propagation as discussed in section 3. 4.1.2 Uncertainties of independen Variables The independent values u and v are each adjusted by controllers [contr]. Their possible deviation from the ideal set point values [sp] u and v depends only on the quality of the control. The maximum possible deviation of u and v from the set points will be denoted as ∆u and ∆v in the following. The deviation is then limited to the control precision. − ≤∆ (9) 272 ∆ (10) The complete possible deviation of the real values μ , and μ , from the set point values is limited by the sum of the possible control variation and measurement error. With the respective overall uncertainties being: ∆ ∆ ∆ (11) ∆ ∆ ∆ (12) 4.1.3 Uncertainties of the dependent Variable Uncertainty of the dependent variable is dependent on the function g: ∆ ∆ ,∆ (13) From the uncertainties of the independent values. Note that function depends on function f and remains to be determined. The measurement of the dependent value itself has also an uncertainty ∆ with overall uncertainty ∆ ∆ ∆ (14) 4.1.4 Comparison of two test stands When applying this methodology, one test stand has to be defined as the reference. The reference measurement is then performed on this test bench. The set points and for the reference measurements could be defined as in sections 2.3. The definition of auxiliary set points follows the central difference scheme in every parameter as given in figure 6: Figure 6: Auxiliary points for comparability study The uncertainty of ∆ , that results from the independent variable is not determined. The basic formula , defined in section 4.1 can be written as f (u , v ) f (u , v ) (15) yiuv u i vi u v v i u u v uv vui i i 273 The partial derivative of the function is not known. However, a reasonable estimate of the slope at the auxiliary points can be gained by applying a suitable numerical differentiation method to the measured data , and . To get sensible results from a numerical differentiation method from measured points, some demands on the performed measurements must be fulfilled. On the one hand, the distance between two auxiliary set points should not be too small, since then the estimation of the slope would be strongly influenced by measurement uncertainties. On the other hand, the auxiliary set points should not differ too much, since the estimation of the slope at the considered specific point can obviously become inaccurate. With the given distance in figure 6 a good compromise is reached. The partial derivation of measurement- and control uncertainty for the operation point results in the following numerical values for the evaluated measurement point: Table 4: Results of partial derivation Reference test Error Control ∆ overall ∆y bench Propagation Uncertainty u= 0.0701 [kg/s] 1.08% 0.23% 0.00092 v = 370.12 [m/s] 0.01% 0.19% 0.74024 y=1.907 [-] 0.21% 0.016 Relative Error 0.86% Two test benches were evaluated following the procedure previously exposed. As Table 4 reports, for the point considered in Figure 6, the reference test bench provide an uncertainty ∆ of ±0.86%. The absolute interval is then 1.72%. This could appear important but all different contributions have to be taken into account. For the same point the second test bench gives a measured value with a 0.72% offset from the reference. The procedure hence facilitates the comparability test. 5 REPEATABILITY The discrepancies which could be considered in a comparability study between two gas test stands could appear too important. Therefore readers could consider the previous paragraphs to determine how many factors to take in account and to get an idea of how discrepancies can be accounted for. To avoid any problem a possible strategy could be to try to perform all the tests that are planned for a development process on the same bench. The test sequence adopted for the testing campaign in this project was specially designed for evaluating the test bench repeatability in reaching exactly the same measurement points. During all test sequence the bench moves between the nine different points of the compressor map, controlling at the same time corrected mass flow-rate and corrected compressor rotational speed. Ten random acquisitions of the same turbocharger working point have been made. Discrepancies are analysed for total-to-total pressure ratio and total-to-total compressor efficiency using the assumption of a perfect gas. Table 5 reports the results concerning the analysed central point of the compressor map, the same as considered in section 3.2. These results are obtained by four different tests performed in a time span of three years. In each column authors report the 2σ interval obtained with following equation (16): 274 ∑ ̅ 2 =2 ̅ (16) The number of points is cut to n, ̅ for the average value and for the different measurements. Table 5: 2σ interval on four different tests on a target point of , = / and , = , / Test , , , , , Continental AG test1 0,08% 0,12% 0,14% 0,27% Continental AG test2 0,19% 0,18% 0,29% 0,23% CRITT M2A test1 0,12% 0,19% 0,20% 0,25% CRITT M2A test2 0,08% 0,23% 0,20% 0,16% Maximum value 0,19% 0,23% 0,29% 0,27% Looking at the results in Table 5, the target mass-flow rate and target corrected speed have two dispersions which are more than half of the final values. For the final calculated variables, not only the combination of these two factors has to be taken in consideration. It is also due to slightly different energy equilibrium in the machine in test. The different equilibrium may be due for example to different heat transfer, test cell and flow enty conditions. As the points are addressed in a random sequence, the influence due to system hysteresis enters the measurement. A big component of this dispersion is definitely due to the accuracy of the control. Steady state parameters were set to 0,3%. Both of the two test facilities gave reproducible results fully in agreement with the stabilisation criteria. Again, as in the case of the representation of the compressor iso-speed line, longer holding time can reduce errors significantly at the price of longer measurement times. 6 IMPROVED TURBOCHARGER TESTING The origins for differences in turbocharger compressor maps are explained in the previous sections. The documentation of the quantitative experiments concerning the differences can be summarized with their principal findings. 1. Users of compressor maps need to be aware of the possible errors in compressor maps. The dispersion due to sensor errors is not negligible. 2. Measurements of a turbocharger compressor map differ due to differences in measurement equipment and boundary conditions. 3. The comparison of two identical test benches with one set of sensors and measurement sections delivers a comparability quality that is acceptable if all the tolerances are taken in consideration. 4. Repeated measurements of one operating point shows only very small deviation in the measurement. Measurements of turbocharger characteristics always have to be made having in mind the usage of the measurement data. A back to back comparison of two different compressors on a single test bench with one set of instrumentation gives precise information on the relative difference. It can be used as a precise tool in turbocharger development and for comparison of two competitor products. To avoid any problem a possible strategy could be to try to perform all the tests planed for a development process on the same bench. 275 ASME. Describing Uncertainties Encountered During Laboratory Turbocharger Compressor tests. S. Mai..-Ing. in flow range or in surge behaviour. H. PTC 10. : Abschlussbericht über ein VFI-Vorhaben. Comparison of the characteristics of different turbochargers measured on different test benches with different instrumentation under different boundary conditions is not sensible. This makes turbocharger test benches an important development tool for turbocharger development if used for comparison tests allowing quantification of for example improvements in efficiency. P. REFERENCES 1.Comparison of measured turbocharger maps and data obtained by computational fluid dynamics (CFD) may differ significantly due to assumed boundary conditions. Mothar. 2001. 2. 5. International. Measurement campaigns for assigned development aims for TC’s should be performed on one dedicated test bench due to significant differences between test benches. 7 SUMMARY AND OUTLOOK Turbocharger maps differ when measurement taken place on a different test bed under different boundary conditions. The difference in boundary conditions in measurement compared to CFD or to engine operation can only be overcome by additional test bench features such as special conditioning units for compressor air.n. Eng. These results should be used with the appropriate assumptions. Measurement Uncertainties Encountered During Gas Turbine Driven Compressor Field Testing. Parameterstudie zur Turbolader- Kennfeldvermessung. The definition of clear performance criteria for measurement could in fact be a useful tool for everyone involved in turbocharger characteristics. R. Gas Turbines Power. Warrendale. s. All that has been said concerns the compressor stage. VDI 2045. SEM Technical Article. Turbocharger Gas Stand Test Code SAEJ1826. and Kurz. Dipl. 3. Repeatability of single operating points on the same test bench with identical instrumentation can be very good. Brun. Chesse. the difference in boundary conditions may lead to significantly different performance maps. Holger. identical inlet and outlet piping compared to the vehicle or other dedicated measurement strategies. ACKNOWLEDGEMENTS Authors would like to thank Continental AG for their assistance in this project. 62-69. 276 .l. 123. J. 4. This would give a complete characterisation of the possible derivations in turbocharger testing. and Chalet. D. When the turbocharger maps are used as input data for engine simulation or engine control unit calibration. The comparison of test results from different test benches with different instrumentation under different boundary conditions may show significant differences. PA : s. This leads to large differences in compressor efficiencies.. Typically CFD computations assume adiabatic walls. March 1995. Comparability of two test benches is acceptable when identical instrumentation and measurement sections are used. The authors plan to analyse the turbine characterisation. K. SAE. and N. Fesich. SAE International SAE 2013-01-0925. 2009. 193-202. M. Kansas : The National Gas Machinery Laboratory. Chapman. 277 . Guidline for Field Testing of Gas Turbine and Centrifugal Compressor Performance. Manhatten. s. 11. S. K. The Efficiency of Turbocharger Compressors with Diabatic Flows. 2013. Turbocharger Heat Transfer Modeling Under Steady and Transient Conditions. Thermodynamische Stoffwerte von feuchter Luft und Verbrennungsgasen. S. Hannover : Dissertation. 7. Chesse and J. Verein Deutscher. Shaaban.F. Experimental investigation and extended simulation of turbocharger non-adiabatic performance. J. Guillou. 2000. 10. 2003. and Nored. Eng. 2004.M. and Casey. 12.l. Hetet. Uncertainty and Measurement Sensitivity of Turbocharger Compressor Gas Stands. 2010. Ingenieure. 8. S. VDI Richtlinie 4670. Brun. Cormerais. Guidelines for Testing large-Bore Engine Turbochargers. J. E. 9. Int. Gas Turbines Power 132. M. P. of Thermodynamics. Schultz. Universität Hannover. M. 2006. Kansas State University. : Gas Machine Research Council and Southern Research Institute.6. K. This paper describes FE simulation methodology development and validation of containment test in order to obtain the optimum housing design at the early design stage. This simulation demands elastic. This paper describes the containment simulation and validation of both compressor and turbine housings. Keywords: Containment simulation. Housing Containment tests verify the operational limits of a turbocharger's housings to contain wheel fragments in the event of a wheel hub burst during operation. explicit FEA. This high speed dynamic simulation methodology involves AWMHS method and explicit FEA. plastic and fracture characteristics. Simulation results are found to be in good agreement with actual containment (rig) tests. ensuring “Design right first time”. S Pandian.Containment simulation and validation of turbocharger housing design J M Ramamoorthy. a wheel hub burst is required. turbocharger housings. S S Parikh. P S Kasthuri Rangan Turbo Energy Limited. area weighted means hoop stress method. ABBREVIATIONS FEA – Finite element analysis AWMHS – Area Weighted Mean Hoops Stress UTS – Ultimate tensile strength KE – Kinetic energy IE – Internal energy TMF – Thermo Mechanical Fatigue ms – milliseconds NOTATION Σaie = sum of element areas σihoop = element hoop stress 1. INTRODUCTION The containment simulation along with TMF simulation helps to verify the optimum housing Design. 2014 281 . For a containment simulation. hardening parameters and ductile damage material model respectively. This methodology is applied for compressor housings and also for high temperature turbine housings. The first part of this paper describes the wheel burst _________________________ © Turbo Energy Limited. wheel burst. which is defined by elastic modulus. ductile damage. India ABSTRACT One of the basic requirements for turbocharger housing design validation is to prevent the wheel fragments from penetrating the housing during wheel burst. If a wheel is considered without material removal. Here a sector model (Figure 2) of the wheel component is considered. material is removed at the bore to get the hub burst effect. In this simulation the wheel is naturally allowed to burst similar to an actual rig test. a weakened wheel is considered for containment simulation. The first part of containment simulation involves ‘wheel-alone’ burst simulation. In wheels.1. AWMHS method Area Weighted Mean Hoops Stress method is a standard method used to estimate the burst speed of rotors. The various phases are explained in detail in the following topics. 282 . The wheel burst simulation initially involves the wheel burst speed estimation. This estimated burst speed is used for explicit simulation. wheel alone is considered (without the housing) to quickly evaluate the wheel weakening dimensions and the material model for wheel burst explicit simulation. As the wheels are designed for a much higher speed margin. 2. various iterations were carried out to arrive at the final design of the wheel so that a wheel hub burst occurs at the required speed. Wheel burst simulation Step 1: AWMHS method Step 2: Wheel burst explicit simulation Housing Containment Simulation Results Comparison with physical testing Figure 1: Workflow for containment simulation and validation 2. This estimated burst speed is used in explicit burst simulation and a material model is developed to match the burst speed. Hence the maximum hoop stress is used to estimate the burst margin.simulation and later the containment simulation and validation. Initially. Using AWMHS method. The hoop stress is estimated using FE Analysis Figure 2: Sector and burst speed margin is estimated using the model of wheel empirical relationship mentioned below. Several iterations are carried out to obtain the optimum material removal dimensions for the required burst speed. The following chart (Figure 1) describes the workflow of the containment simulation & validation. Since the wheels are designed with a good speed margin. In the case of wheels under rotating conditions hoop stress is of major concern. The wheel burst is initially estimated using AWMHS method. WHEEL BURST SIMULATION A wheel bursts when the centrifugal force exceeds its internal binding forces. it may lead to over design of the housings. it is required to weaken the wheel. This high speed and short time duration simulation can be performed only in explicit analysis. The uniqueness of this project is that the wheel is allowed to fail by itself in simulation and that no restriction is imposed on the no of failed segments. 283 . Wheel burst explicit simulation The turbocharger wheels rotate at high speeds and the wheel burst occurs over a short duration of time. ∑ 𝑎𝑒𝑖 ∗ 𝜎ℎ𝑜𝑜𝑝 𝑖 𝐴𝑊𝑀𝐻𝑆 = ∑ 𝑎𝑒𝑖 𝐵𝑢𝑟𝑠𝑡 𝑆𝑝𝑒𝑒𝑑 𝑀𝑎𝑟𝑔𝑖𝑛 = √(𝑈𝑇𝑆⁄𝐴𝑊𝑀𝐻𝑆) The workflow to arrive at the weakened wheel design for containment simulation from AWMHS method is shown in Figure 3. For turbine wheels higher temperature material properties are used. a material model for explicit wheel burst simulation is chosen.2. Different regions of the stress-strain curve specific to this project are shown in Figure 4[1] [2]. Cyclic sector wheel model FE structural analysis Modify material removal dimensions Calculation of burst speed Check if burst No speed = required speed Yes Conclude Figure 3: Workflow to obtain material removal dimensions using AWMHS method 2. Since the failure involves fracture of the component. The burst speed and weakening dimensions obtained from AWMHS method are used herein as input. The elastic region involves the elastic modulus. plastic and also the fracture characteristics for the material are to be defined. elastic. In this project based on the burst speed obtained from AWMHS method. the plastic region involves plastic strain and the fracture region requires the damage criteria which include the failure strain. For turbine wheels high temperature material models are used. thereby reducing the time and cost for extensive material testing. thus a model to capture its ductile nature is used. The wheels have an elongation of greater than 5%. The damage criteria in the material model were developed iteratively to match with the burst speed obtained from AWMHS method. Plastic σ Elastic Fracture ε Figure 4: Different regions of stress-strain curve There are various damage models available in commercial software [3] [6]. The workflow to obtain damage criteria for explicit wheel burst simulation is shown in Figure 5. The damage model is very critical as it defines the failure of the wheel. similar to failure of wheel in actual operating condition. Input: Weakened geometry & burst speed from AWMHS Damage criteria Explicit burst simulation Modify Damage criteria Burst @ Required No speed Yes Yes Burst @ 1% < Required speed No Conclude Figure 5: Workflow to obtain Damage Criteria 284 . in the event of a wheel burst. temperature dependent material models are used as the temperature on the turbine housing is non-uniform.1 ms 0. in the early design stage. They also act as a safety standard for the turbocharger. A strain rate dependent material property is required as the impact velocities of wheel fragments on housing are not constant. 3. The criteria for a containment test are that. for turbine housings. These results are for a time period of 0. to visualize the actual damage on the housings in simulation. HOUSING CONTAINMENT SIMULATION Containment test along with TMF test decide the design criteria for the housings. Hence an upfront containment simulation. Figure 6 shows compressor and turbine wheel burst in explicit simulation. a suitable damage criterion is developed using the AWMHS method. housing and central housing cover are considered.16 ms 0. Demonstration of this requirement is an expensive and time-consuming process. no fragments of the wheels must penetrate the housings [5]. For containment simulation. (a) (b) Figure 6: Wheel burst explicit simulation (a) compressor wheel (b) turbine wheel 3.Thus. For compressor side temperature effects are not considered.3 ms. both the wheel and housing are considered which makes the problem larger in size and more complex.18 ms 0. The damage on the compressor housing can be visualized in Figure 7. Additionally. 0. Compressor containment simulation In this simulation compressor wheel. By following this approach an optimum and robust design of the housing can be arrived at.14 ms 0. Material models for the housings are developed.3 ms Figure 7: Compressor housing containment 285 . would lead to substantial time and cost savings.1.24 ms 0. similar to that of wheels. To include the temperature effects on turbine side a steady state thermal analysis is performed for the housing and the temperature distribution is mapped on to it in containment simulation.4 ms 0.The total duration of the explicit simulation must ensure considerable reduction of kinetic energy of the wheel [4]. The damage on the turbine housing can be visualized in Figure 9. This can be confirmed by plotting the energy graphs as shown in Figure 8.6 ms.2 ms 0. These results are for a time period of 0. 0. Turbine containment simulation In this simulation turbine wheel and housing are considered.1 ms 0.5 ms 0.6 ms Figure 9: Turbine housing containment 286 . KE of compressor wheel = IE of Compressor wheel + IE of Compressor housing + IE of Central Housing Cover + Friction Losses Figure 8: Energy graph for compressor containment simulation 3.2.3 ms 0. The energy graph for turbine side is shown in Figure 10. The Figure 11 shows the results comparison of compressor containment rig test and simulation. RESULT COMPARISON WITH RIG TEST The simulation results are compared with several rig tests and validated. Figure 11: Compressor containment result comparison 287 . A good agreement is seen between Simulation and Tests on the pattern of ‘two-parts burst’ at same location observed in both rig test and simulation. In both compressor & turbine side the wheel burst speed Simulation and rig test results towards the Burst speeds is seen to be less than 2%. The housing contains the wheel burst in both rig test and simulation and no penetration observed. This validates our material model developed using AWMHS method and wheel burst explicit simulation. Normalized Internal Energy KE of Turbine wheel = IE of Turbine wheel + IE of Turbine housing + Friction Losses Figure 10: Energy graph for Turbine containment simulation 4. 11 Documentation. 2012 4. José L. 2004 Abaqus users conference 288 .The Figure 12 shows the results comparison of turbine containment rig test and simulation. Abaqus explicit advanced Training manual. This simulation methodology along with TMF simulation (for turbine housings) helps to design housings with optimum thickness in the early design stage.. Figure 12: Turbine containment result comparison 5. 2.0 version. 2011 3. The housing contains the wheel burst in both rig test and simulation and no penetration observed. Generally the thickness of housings is decided with a high factor of safety. Garret by Honeywell 6. 5. Proceedings of the IMPLAST 2010 Conference October 12-14 2010 Providence. White paper No. to prevent containment failure in rig test. Alcaraz. © Dassault Systèmes Simulia Corp. “Blade Impact Simulation Against Turbine Casings”. Copyright © 1990-2011. Optimum thickness of the housings is also likely to improve the life of the housings considerably. Thus the simulation results are validated with that of rig tests. Similar pattern of hub burst at same location is observed in both rig test and simulation. This leads to bulky housings and also in case of turbine housings. The housings contribute to more than 60% of the total weight of the turbocharger. REFERENCES 1. which fulfils the need of the hour. Analysis user’s manual.. Rajeev Jain. it reduces the TMF life of the same. Radioss Theory Manual 11. 6. Burst & containment : Ensuring Turbocharger safety. © Dassault Systèmes. Rhode Island USA © 2010 Society for Experimental Mechanics. Abaqus 6. “Prediction of Transient Loads and Perforation of Engine Casting During Blade-Off Event Fan Rotor Assembly”. Unai Hermosilla. Substantial material savings were observed by implementing this simulation methodology in housing design. Aja. Inc. It is also helpful to identify unwanted material addition in the housing design. Altair Engineering. Inc. CONCLUSION Today the OEM market demands lighter turbochargers with high durability. Angel M.2. A turbocharger equipped with the detection system has been adapted to a modern four cylinder gasoline engine with direct injection. In the paper at hand. Hence. In this context. High quality turbocharger performance data are necessary over a wide range of operation conditions as input for engine simulation programs. on-engine performance of the turbine stage. Complete drive cycle simulations (e. M Bargende2 1 IHI Charging Systems International GmbH. the volute and/or the wheel behave quasi-steadily or have to be considered as unsteady devices. _______________________________________ © The author(s) and/or their employer(s). D Filsinger1. Engineering Division. Furthermore. the results give a clear indication about the significance and magnitude of unsteady effects within the turbine stage under pulsed flow conditions. Institut für Verbrennungsmotoren und Kraftfahrwesen. The available measurements represent an excellent basis for advancements in the modelling and simulation of turbocharger turbine stages with engine simulation tools. Especially modelling the turbine stage efficiency for engine-like operating conditions (pulsed flow) still is under research. It is possible to measure the turbine shaft torque with high accuracy and time resolution.g. a contactless shaft torque detection technique . Germany ABSTRACT The trend to use advanced simulation tools for engine performance prediction is continuing and even emphasized due to shortening of development cycles.is presented. many researchers raised the question about unsteady effects within the turbine stage and whether the stage. the accuracy of turbine and compressor maps becomes more and more relevant to achieve reliable simulation results and predictions. The turbocharger has recently developed away from an auxiliary part towards an integral component of the internal combustion engine. The highly accurate prediction of steady and transient engine behaviour becomes increasingly important. Germany 2 Universität Stuttgart. 2014 301 .that has been integrated into an automotive turbocharger .Engine crank angle resolved turbocharger turbine performance measurements by contactless shaft torque detection B Lüddecke1. NEDC) help to assess turbocharged engine performance at very early stages of complex engine and vehicle development projects. Engine cycle resolved torque data has been gathered in order to assess the crank angle resolved. 5 turbine impeller inlet 4 turbine stage exit aero corresponding to aerodynamics air / a values related to air C compressor (stage) Cycle related to the full engine cycle coolant state of coolant corr corrected eff/effective relevant/operative part fric/friction related to friction gas / g values related to burnt gas GE related to gas exchange i crank angle index during engine cycle. blade speed ratio [-] Abbreviations BDC Bottom Dead Centre CA Crank Angle CFD Computational Fluid Dynamics DFT Discrete Fourier Transformation ECU Engine Control Unit MFP Mass Flow Parameter [m s K^0.NOMENCLATURE A Area [m^2] ∆t time difference k kilo. 1. [°C] for tables) u blade tip speed (turbine or compressor) [m/s] .5] NEDC New European Driving Cycle TDC Top Dead Centre WOT Wide Open Throttle Greek Symbols ∆ difference [-] number pi ɵ moment of inertia [kg m^2] Engine cycle duration (720 °CA) Indices/Subscripts 1 compressor stage inlet 2 compressor stage outlet 3 turbine stage inlet 3.000 T torque [Nm] mass flow rate [kg/s] n rotational speed [rpm] P power [W] PI pressure ratio [-] p pressure [Pa] T temperature ([K] for calculations. i ∈ {0-719} inertia related to rotor inertia is isentropic journal related to journal bearing lubricant state of lubricant meas measured value r radial 302 . the volute and/or the wheel behave quasi-steadily or have to be treated unsteadily. In the present work. the results of a detailed sensitivity study regarding sensor/rotor relative displacement are presented. Furthermore. mass flows and rotor speeds. on-engine performance of the turbine stage. One main advantage of this approach is its robustness towards occurring heat flows into the compressor stage during testing (compare [2]). Furthermore. Engine cycle resolved torque data has been gathered in order to assess the crank angle resolved. the results give a clear indication about the significance and magnitude of unsteady effects within the turbine stage under pulsed flow conditions. 2 SENSITIVITY STUDIES The influences of varying distances between the primary sensor (magnetically coded shaft) and the secondary sensor (coils + electronics) were investigated to prove the robustness of the system and its suitability for the conducted investigations. The available measurements represent an excellent basis for advancements in the modelling and simulation of turbocharger turbine stages with engine simulation tools. However. turbine map extrapolation methods are state of the art.rotor corresponding to the rotor s static ss static-to-static T turbine (stage) t total tt total-to-total ts total-to-static 1 INTRODUCTION In [1] a novel. based on inverse magnetostriction was introduced. In this context many researchers raised the question about unsteady effects within the turbine stage and whether the stage. The dashed grey lines indicate the maximum potentially possible movement that could theoretically occur considering geometrical clearances and tolerances of the bearing design. Furthermore. the detection system is employed for unsteady torque measurements under pulsed flow conditions on an internal combustion engine. The direction of the investigated movement is indicated by the white arrows. Especially the modelling of turbine stage efficiency for engine-like operating conditions (pulsed flow) still is under research. however they still suffer from non-perfect input (test) data. Of course. contactless turbine shaft torque detection system. the relative influence of bearing losses on turbine performance measurement is reduced significantly. It was employed for steady state performance measurements of a mixed flow turbine on a hot gas stand. In Figure 1 the rotor and sensor of the torque sensing system are shown in the lower right corner. 303 . the occurring displacements and movements of the rotor under real conditions are usually significantly smaller due to oil film thickness and the damping effects of the oil film onto shaft motion. This enables a reasonable turbine performance measurement especially at low pressure ratios. The standard serial production turbocharger of a modern four cylinder gasoline engine with direct injection has been equipped with the shaft torque detection. Consequently. During expansion.as the differential signal of the coils is used to record changes in the outer stray magnetic field of the primary sensor . where many residual sources of electromagnetic fields exist.even under worst case scenario assumptions.the system is very robust against parasitic magnetics fields (like geomagnetic fields or other sources).Figure 1: System configuration and results of sensitivity studies regarding rotor/sensor relative displacement Hence the influence of the rotor motion in any direction becomes quite negligible within the range that the rotor can potentially move during operation . This is referred to as the aerodynamically available torque. e. The torque sensor is located between the journal bearings of the turbocharger. This robustness becomes especially important at engine test stand. The minor losses by seal ring friction as well as the ventilation losses on surfaces are not considered. the ignition system of the engine. The change of an external magnetic field will influence the absolute signal level of every coil.g. the measured value is lower than the actual aerodynamic torque. Figure 2 illustrates how the torque is transferred from the turbine to the compressor through the shaft for steady state conditions (no rotor acceleration or deceleration). as the turbine side journal bearing causes a torque loss. while the difference between them remains constant. Furthermore . as described in [1]. The 304 . 3 SHAFT TORQUE AND BEARING FRICTION To better understand the relevance and interpretation of the measured torque data. the gas flow causes pressure forces on the blade surfaces and hence transfers torque to the impeller. eq. 200. to keep the thermal loading for the measuring system low. 400. As mentioned above. which significantly differs from the steady case. It is important to draw the attention to these relations in detail. resulting in pressure rise. Figure 2: Schematic of torque transfer from turbine to compressor and occurring losses for steady state condition 4 HOT GAS STAND MEASUREMENTS Prior to engine testing. . 305 . as these losses are modelled and corrected based on available bearing friction measurement data. the torque is measured between the two radial bearings. Hence. this has no implication on the relevance and accuracy of the data. a polynomial function (of the rotor speed) has been employed to account for the journal friction torque. of course. 800. However. as below in chapter 7 the unsteady case under pulsed flow (with alternating rotor speeds) is considered. The supply pressure of the lubrication was set to a constant value of 4 bars absolute. to assess the corresponding aerodynamic torque – that is basically the one of interest – the friction torque of one radial bearing has to be added to the measured torque. In Figure 3 an excerpt of the measured turbine torque for T3=600°C is plotted against the total to static pressure ratio. (1). This leads to higher values of bearing friction. the correction causes an offset from the measured value. a detailed experimental study on the hot gas stand was elaborated under steady state boundary conditions. respectively. The turbine inlet temperature T3 was varied in six steps (40. = + (1) Based on available experimental data. 600. The temperature of coolant and lubrication were set to 30°C and 35°C. Figure 2. 1000°C) and for every case several speed lines have been recorded. The result is given by the black curves with white circles in Figure 3. Obviously. The remaining torque is available to the compressor wheel for work input into the fluid.compressor side journal bearing as well as the thrust bearing create additional losses. The red curves with squared symbols represent the measured “raw” torque. while the slope remains constant. The white circles accordingly mark the location of the numerical results that the plot is based on. the corresponding turbine stage efficiency – based on measured shaft torque . (1)) as demonstrated in Figure 3 and hence represents aerodynamic efficiency. Figure 4.is plotted. pressure ratio Additionally. bottom. This is to be expected and has also been reported by several researchers. for every speed line. [4]. compare e. The CFD data also show a fairly linear trend but as the computed range is much wider than the measured one. The white circles represent exactly the data points from the plot at the top. corresponding to the turbine inlet temperature. the measured data and the CFD results in particular show that the turbine rotor torque would become zero far before the pressure ratio equals one. The comparison between the (bearing friction corrected) measured data and CFD results shows good agreement. [3]. the measured data show a linear dependency of turbine rotor torque versus stage pressure ratio within the measured range.g. friction corrected and computed data match well. a contour plot based on CFD data is shown. The reason for this is believed to be related to the strongly changing flow regimes and mach numbers. [5]. This is related to the centrifugal force field of the rotor (compare [6]) and is important for the interpretation of the engine results shown later. 306 . The available recorded data for different rotor speeds as well as turbine inlet temperatures have been employed to gain a wide map of the turbine stage in terms of blade speed ratio and pressure ratio. At Figure 4. torque data from CFD results are shown as solid blue curves. top. if for constant rotor speed the pressure ratio is varied in a wide range. In the middle of Figure 4. if very low and very high pressure ratios are compared. which have been used to interpolate a contour of turbine stage efficiency. The iso-lines have a distance of one per cent efficiency. it can be stated that measured. Obviously. The plot shows the efficiency after journal bearing friction correction (eq. The absolute level of efficiency though is still highest for the computed data. However. smooth changes in the slope can be seen. Furthermore. Figure 3: Measured and calculated torque trend vs. The data points of the measured speed lines are marked with different symbols. Compressor and turbine specification were not changed. incorporating the described contactless shaft torque detection system. These are for instance: geometry simplifications and turbulence modelling for the numerical results. 5 ENGINE MEASUREMENT SETUP AND INVESTIGATED OPERATING POINTS Based on the serial production turbocharger specification of the test engine. It is also conceivable that the radial bearing friction model does not perfectly represent the occurring bearing friction for all hot gas temperatures and all speed lines as the model is based on a limited set of experimental data. the data shown above give a very high level of confidence into the presented method for steady state conditions. regarding engine-turbocharger matching. deviation of real geometry from nominal shape as well as the inevitable uncertainties within an experimental data set. The standard serial production turbocharger was then replaced by the prototype. 307 . Hence.Figure 4: Turbine stage efficiency contour plots based on measured torque and radial bearing friction corrected (top) and comparison with CFD data (bottom) This can be attributed to miscellaneous minor and partly superimposed influences. the engine can be operated safely within its full operating range. a prototype turbocharger was built up. However. For the time resolved turbocharger turbine shaft torque measurements.1 °CA resolution). rpm % % °CRK °CRK -- 116_00 1250 WOT / 100% 63.18 It is obvious how significantly the valve overlap timing influences engine torque.6 82 -80 1. employing the commercially available software Tiger [10]. Speed Pedal Load Cam_int Cam_exh Lambda -. the spectrum of the measured signal was investigated regarding its bandwidth and maximum relevant frequency. The OEM follows a conservative scavenging strategy to assure durability of the engine as well as the definite avoidance of pre-ignitions under any circumstances in the field. This workflow assures high quality crank angle resolved data and reasonable final file sizes.9 82 -72 1. The valve effective 308 . which was used to arrange special parameter variations for detailed investigations of the interaction between the combustion engine and the charging system. hence the limitations mentioned above can be exceeded. For the investigated case the engine had been well conditioned and was operated under surveillance of a monitoring and control system. By this well- known operating strategy ([7].6 110 -110 0. Although – due to the complex flow regimes – scavenging operation is hard to analyse accurately by a zero-dimensional or limited one-dimensional code.99 116_01 1250 WOT / 100% 78. engine torque can be almost doubled. The load is given as percentage. All points have been recorded at an engine speed of 1250 rpm. post-processed. a recording frequency of over 100 kHz was used. The increased mass flow over the turbine wheel leads to increased turbine shaft power and hence compressor power. Four stationary stable operation points are presented. In Table 1 an overview of the engine operating points presented in this paper is given. what is believed to be related to two main reasons: The waste gate had been mechanically blocked to minimize leakage – a state that can surely not be reached in serial production engine under pulsed hot gas conditions. [8]. [9]). the results clearly indicate the fraction of scavenged mass.g.09 116_03 1250 WOT / 100% 112. as a partly programmable ECU was available. Any engine load variation is solely caused by intake and exhaust valve timing variation and thus strongly related to the so called “scavenging” mechanism that (additionally to the conventional turbocharging) boosts the engine. where only the cam phasing for intake and exhaust valves was varied to adjust engine load. a combined combustion and gas exchange analysis was conducted exemplarily for cylinder four. For the shown steady operating points. certain conditions of operating points could be manipulated independently. filtered and then an average engine cycle was calculated. By DFT. Even the serial application full load torque can be exceeded. For the operating points listed in Table 1. e. while the throttle was kept at WOT conditions. related to the full load torque of the serial production engine application at 1250 rpm. The corresponding results are plotted in Figure 5. valve timing. The raw data were then filtered appropriately and resampled to crank angle (0.Furthermore.5 85 -85 1 116_02 1250 WOT / 100% 91. All four stationary points were run at thermal persistency and close to engine knock limit. Table 1: Engine operating points No. about 200 consecutive engine cycles were recorded. as for high speeds and loads the longer duration is needed (together with gas dynamic effects) to realize the exchange of the cylinder gas mass in a very limited period of time. as intake and exhaust valves are allowed to be opened simultaneously with a certain overlap. However. This is also one main driver for twin scroll or double scroll concepts. For scavenging. The intake and exhaust channel pressures of cylinder four are drawn as solid blue and red curves. At the end of the scavenging process (close to exhaust valve closing). a short exhaust valve opening duration can only be used effectively for the low end torque area. This effect proves the imperfect flow separation of the exhaust channels in especially four-cylinder engines. it enables scavenging at all. The corresponding calculated intake and exhaust mass flows are drawn as dotted blue and red curves. cylinder inlet pressure (~compressor outlet pressure) has to be higher than cylinder outlet pressure (~turbine inlet pressure). [8]. An alternative is a variable exhaust valve opening event. In a four cylinder engine. a negative mass flow can be observed. a shortened and/or variable exhaust timing length can help to avoid this. where the exhaust valve opening duration is longer than the distance between two exhaust strokes. Secondly it also moves the relative position of pressure pulses and valve overlap in the desired direction. where the flow separation is realized within the turbine housing.areas are given by the dashed black lines. realizing this flow separation within the cylinder head. 309 . as shown in [7]. the shift of the valve lift curves causes two changes: First of all. Figure 5: Results of gas exchange analysis Obviously. This is caused by the absolute length of the (exhaust) valve timing as the opening event of the next cylinder’s exhaust valve pushes back some mass flow while the exhaust valve of the actual cylinder is still opened. 9 29 1.5 °C bar abs °C -. -. The same is true for the right graph.is shown and the four compressor operating points are depicted. the waste gate was mechanically blocked to assure that the complete gas mass flow is fed to the turbine wheel.49 1. Moreover. the turbocharger conditions corresponding to the engine operating points of Table 1 are listed. due to packaging. Distinct colours are used for each operating point within the graphs (116_00=blue. 116_03=black/white). 116_01=green. PI_C PI_T Turbine lubricant lubricant coolant AVG position rpm K^-0.9 29 1. Table 2: Turbocharger operation points corresponding to Table 1 N_red T_ p_ T_ T_3 Wastegate No.18 1. The avoidance of any significant leakages (internally or to the test cell environment) is mandatory to achieve reliable power balances.34 722 blocked 116_03 3773 25 3. Although measurement positions. signal time averaged). as the average turbine and compressor pressure ratios increase.85 1. 310 .30 1. The four single data points are based on averaged measured pressures (fast sensors. allowing for sensibility crosschecks of the measured data. Figure 6: Graphical overview of investigated operating points: compressor map (left) and turbine map (right) In the left graph the compressor map – obtained from hot gas stand tests .9 29 1. In Figure 6 an overview for 116_00 to 116_03 (Table 1 & Table 2) is given.8 29 1.23 689 mech. showing the steady state hot gas stand turbine map (T3=600°C) and averaged turbine data during engine operation. piping and bends are different on engine compared to the gas stand measurement.15 649 closed / 116_01 2454 23 3. time averaged signal) from engine test stand. temperatures (slow sensor principle.57 758 The turbocharger speed develops as expected for increased engine load. In Table 2. 116_02 2996 24 3. 116_02=red. the actual position of the waste gate system was controlled by an optical distance sensor to prove the blocking forces were high enough to resist the opening forces caused by the exhaust gas pressure pulses acting on the waste gate surface. a good match is obvious.6 CYCLE AVERAGED ENGINE AND TURBOCHARGER OPERATION POINTS AND TOTAL VALUE CORRECTION As mentioned above. no averaging) and turbocharger rotor speeds (fast. °C -- 116_00 1970 22 3. there is no straightforward method available to extract the total values from the available measured data. For the compressor. The phase shift is depending on the volute size and design. with the difference that at the turbine inlet the variations/pulsations are much larger and the temperature. respectively. No correction is computed as it has been done for the presented steady hot gas stand maps. Many authors raised the question how to correctly define and calculate a stage (and/or wheel) pressure ratio that is comparable to steady hot gas stand results. [14]. The results are presented at the end of this paper in chapter 8 after all other relevant phenomena have been discussed and corrections have been introduced. Hence it was decided to conduct a sensitivity study by applying phase shifts onto the measured inlet and outlet pressures. [5]. turbine stage pressure ratio and turbine rotor torque 311 . the pressure tapping positions as well as the engine operating point with its rapidly changing flow and temperature regimes during the engine cycle. [11]. The potential error introduced by a possibly not fully valid correction procedure could be higher than the potential gain in accuracy. Especially during blow down phase (shortly after exhaust valve opening) the flow velocities and temperatures as well as their gradients raise significantly. Figure 7: Crank angle resolved pressures. due to the complex packaging situation and especially because of the unsteady flow conditions (consecutive filling of the cylinders). impeller inlet and turbine stage outlet is a frequently discussed topic when analysing time or crank angle resolved turbine performance data. the difference between static and total values amounts to only few thousandth of a bar.7 ENGINE CYCLE RESOLVED TURBOCHARGER AND ENGINE MEASUREMENT RESULTS The phase shift between turbine stage inlet.and velocity level is higher. It is to note that all cycle resolved pressures and pressure ratios from engine measurements are static and static-to-static. Anyway. The same is true for the turbine side. no straight line with constant slope. As the measured. the employed averaging procedure is very important in terms of achieving results that are comparable to the steady case. coloured numbers for unsteady results).In Figure 7 the crank angle resolved pressure data for turbine inlet (red solid line) and outlet (blue solid line). while the comparison with turbine inlet pressure pulses exhibits a shift in time. all values plotted vs. As mentioned by other authors ([3]. Figure 8: Measured unsteady shaft torque. The unsteady data is cycling in the vicinity of the steady state measurement data. are shown. In this context. average turbine inlet temperature during engine operation is in between these two values. energy averaged values and steady-state torque data obtained at hot gas stand. however. but an orbit is observed. the steady torque data (also bearing friction corrected) obtained at hot gas stand for T3=600°C (grey squares) and T3=800°C (grey triangles). A clear dependency of measured turbine torque and stage pressure ratio can be seen. [5]) and as expected. The corresponding reduced turbine speeds are given as well (grey numbers for steady state. It is obvious how the amplitudes and gradients of all plotted values increase with engine load and that due to valve timing variation also the location of the pulses changes. In Figure 8 the measured. The values have already been corrected for the journal bearing friction torque. turbine stage pressure ratio As a third group of values. The averaged values shall represent a process that transfers the 312 . Compared to the usually presented MFP values (comparison of steady and unsteady case. [12]). this is the best comparison that can be made between pulsed and steady flow conditions. see [3]. the energy averaged torque and pressure ratio are plotted as coloured asterisk symbols. due to filling and emptying effects of especially the volute. Additionally. compressor outlet (green solid line) as well as the stage pressure ratio (black dashed line) and the rotor torque (black solid line) are shown for all four operating points described in Table 1 and Table 2. unsteady turbine rotor torque is plotted versus turbine stage static-to-static pressure ratio. the deviations as well as the observed orbit are quite small. [11]. The reason for this unexpected slope is the transient speed change of the rotor. (2). the overall amount of energy that is transferred from the turbine wheel over the shaft to the compressor stage (and the bearing system) during one engine cycle is calculated. eqs. Some minor deviations exist. [6]. This could be related to unsteady turbine operation. the average turbine shaft power can be computed. = (5) . One also observes a difference in the slopes for unsteady and steady state measurements in Figure 8. relevant cycle average turbine stage pressure ratio can be determined by eq. Figure 9 shows the measured stage pressure ratios as already shown in Figure 7. For the steady state results it is clear that the torque will become zero before the pressure ratio equals unity due to the centrifugal force effects of the rotor. hence the asterisk symbols do not match exactly with the (extended. = 2 ∆ = ∆ (2) 60 The value ∆ is not quite constant during engine cycle. As a matter of principle the measured temperatures behave more like a time averaged quantity and hence do not represent energy equivalent values. = = ∑ ∆ Together with the cycle average turbine shaft speed. . . . (6). 313 . First of all. . However. However. eq. Finally. eq. . . ∑ ∆ It is obvious that the cycle-average values of torque and pressure ratio are close to the steady state speed lines for T3=600°C and T3=800°C. . the energy equivalent. ∆ . ∑ . it seems more likely that the “slow” temperature measurement method and its positioning are of major importance in this context. Based on the cycle energy and the engine cycle duration (720°CA). (3). ∆ = (6) . This characteristic is not exhibited for the measured unsteady torque data. they are almost in between or in extension of the steady measurements with comparable reduced speed. as small variations in engine speed can be observed. ∑ ∆ (3) . cycle averaged shaft torque is calculated. 2 . if necessary) reduced speed lines at the energy averaged pressure ratio.same average power to the compressor stage (and the bearing system) as a steady state operation on hot gas stand would exhibit. More precisely. the energy equivalent. =2 = (4) 60 60 ∑ ∆ . Please note that the journal bearing friction corrected torque values are employed. (4) & (5). this is an aerodynamic matter of fact that should not change for radial turbomachinery under any condition. Thus it is addressed with the index “effective”. according to eq. for acceleration as well as for deceleration. = + + ∙ 60 ∙ 2 (8) . . 314 . eq. (8). related to rotor acceleration/deceleration. The correct value of (aerodynamic) torque can be determined by superimposing both torques as well as the bearing friction correction for the radial journal bearing. rotor torque and rotor speed amended by the calculated torque based on speed variation and inertia The crank angle resolved characteristics of the inertia based torque look similar to the contactlessly measured values with the difference that a change in sign occurs. . Figure 9: Measured pressure ratio. The practical interpretation is that during the acceleration phase.amended by the measured rotor speeds as well as the torque values that are based on the rotor speed change and the rotor inertia. For calculation of this torque. not the full inertia of the rotor. The described behaviour is illustrated in Figure 10. It is obvious that the rotor speed change corresponds to a significant amount of torque (in the order of magnitude of the measured signal) and hence has relevant influence on the measured shaft torque and certainly cannot be neglected. only the inertia of the turbine wheel and that portion of the shaft until the measuring position has to be taken into account. (7). = ∙ 60 ∙ 2 (7) . the rotor “swallows” some of the aerodynamic torque (hence it can’t be measured) and releases it again during deceleration (hence more than the aerodynamic torque is measured). especially for low pressure ratios. For the upper right graph in Figure 11. where the turbine stage performance has been investigated under pulsating conditions by the use of a special test rig and instantaneous measurements. 315 . respectively. An energy equivalent value that shall be compared with steady state values from hot gas stand only makes sense. if it based on the (measured) shaft power that is in fact transferred to the compressor. (7) is given. In the upper right corner. For high pressure ratios. Regarding the (energy) averaged values (asterisk symbols). In the upper left graph the measured torque during steady and unsteady operation for case 116_03 is shown (compare Figure 8). While measured torque shows invariably positive values. The superimposed values look much more reasonable than before correction. the calculated torque. acceleration and offset influence onto shaft torque measurements In Figure 11 the results before and after inertia torque correction are shown again versus static-to-static pressure ratio. Hence it is not sensible to calculate an energy average pressure ratio that belongs to the blue orbit. a very good agreement is obvious. The lower graph shows the superimposed curves of the two upper graphs. the inertia based torque can be positive or negative. as no net energy is effectively conducted via the shaft to the compressor. the orbit has grown and henceforth encases the steady state values from hot gas stand. an important note is to make. based on eq. The extrapolated curve would attach zero torque before pressure ratio equals one. according to acceleration and deceleration. Figure 10: Rotor shaft torque distribution for unsteady case: deceleration. If the slope of the resulting curve is compared with the steady values.An equivalent approach has been presented in [13]. as anticipated. the transferred energy portion is zero (apart from decimal places). however. all values plotted vs.analogous to the method from Figure 11 – are plotted. the slopes of the torque orbits are much more reasonable and for low pressure ratios show good match with the steady state data (grey squares and triangles). energy averaged values and steady-state torque data obtained at hot gas stand. Figure 12: Superimposed torque curves (measured + inertia based). Figure 11: Measured torque after bearing friction and inertia correction In Figure 12 the correction results of all curves from Figure 8 . turbine stage pressure ratio 316 . Obviously. rotor inlet/outlet and stage outlet has not been considered. 8 SENSITIVITY STUDY REGARDING PHASE SHIFT CORRECTION As explained. as recommended by other researchers.3 2.9 116_03 10. Hence.6 2. an assumption that is supported by the findings in [14] for an automotive size turbocharger turbine stage. Table 3.2 11. a proper (and even cycle resolved) phase shift correction is hard to calculate analytically. Gas flow Speed of Super- velocity sound imposition velocity sound imposition °CA °CA °CA °CA °CA °CA 116_00 20. It was necessary to discuss the steady and unsteady turbine stage characteristics as well as the corrections that have been introduced above. half of its circumferential length has been employed.2 1 0. After correction of all relevant influences any orbit should then collapse to a line and this line should be congruent to a measured steady state speed line obtained at the hot gas stand.6 2. taking into account gas flow velocity. As characteristic length within the volute geometry.2 14. speed of sound or a superimposition of both. In Figure 13 also the steady state values from hot gas stand that fit approximately in terms of average reduced speed are plotted additionally (blue squares and triangles). it is not claimed that the presented phase shift results represent a perfect reproduction of real flow physics.4 19. speed of sound and their superimposition Phase shift correction study results for Phase shift correction study for turbine turbine stage inlet pressure p3 (delay) stage outlet pressure p4 (advancement) Gas flow Speed of Super.9 It is believed that the most reasonable approach takes into account both gas flow velocity and speed of sound and that p3 is delayed while p4 needs to be advanced. The result of this variation is plotted in Figure 13. that the rotor behaves quasi-steadily and the orbits are caused by volute filling and emptying effects.1 2. 317 . This is to be expected if the fundamental hypothesis is. [14]. A sensitivity study was undertaken for a phase shift of p3 and/or p4 solely as well as simultaneously. temperatures and pressures. Obviously. the phase shift between stage inlet. However.g.So far.4 1 0. Especially for the low and mid pressure ratios a very good match for the slopes and absolute values is obvious. prior to addressing this topic. average measured mass flows.7 2. as already mentioned in chapter 7.2 2.2 2. e.1 1 116_01 16. Table 3: Results of phase shift correction study for measured turbine stage inlet and outlet pressure considering gas flow velocity.9 1 1 116_02 13. the gas flow velocity as well as the speed of sound was estimated by compressible flow equations.7 1. the observed orbits after phase shift (red solid curves) become significantly smaller compared to the cases without phase shift (black dotted curves). based on the known nominal geometry.5 2 8. expected relationship between pressure ratio and rotor torque. turbine stage pressure ratio before (dotted black curves) and after (solid red curves) phase shift (parallel delay of p3 and advancement of p4) and comparison with steady-state data The remaining orbit surface areas (discrepancy between rising and falling pressure ratio limb of the unsteady torque measurements) can be explained by: Remaining. not fully captured volute filling and emptying effects (due to simple correction procedure with gas flow velocity and speed of sound. Fractions of the rotor speed change that are related to an alternating compressor mass flow (and hence its power consumption) caused by the intake strokes of the four cylinders and not related to turbine power fluctuation. This influences the result of eqs. (7) and (8). the turbine torque trend exhibits almost steady characteristics. Real unsteady effects in the turbine wheel / volute-wheel interaction for very high pressure gradients as well as mass flow gradients. Figure 13: Torque vs. especially for low stage pressure ratios. especially since being based on average measured values and not cycle resolved). Scavenged mass flow during gas exchange at valve overlap timing that alters the usual. Engine speed variations during one complete cycle. “Unsteady” / not captured effects due to the rapid change of reduced speed / highly instantaneous temperature fluctuations shortly after engine exhaust valve opening. However. 318 . it is to note that after phase shift correction. No consideration of transiently changing degrees of reaction during the pulsed flow admission. A. Seiler. F. Filsinger. L. R. T.. San Antonio. GT2013-95815 319 . the role of phase shift correction was discussed exemplarily. Martinez-Botas. of ASME Turbo Expo 2013: Power for Land. 10th International Conference on Turbochargers and Turbocharging of the IMechE. USA. C. Sea and Air.. Proc. [3] Cao. the relevance of bearing friction and especially rotor speed change was pointed out.. M.. (2013) “A Method of Map Extrapolation for Unequal and Partial Admission in a Double Entry Turbine”. C. REFERENCE LIST [1] Lüddecke. Sea and Air. ASME Turbo Expo 2013: Power for Land. M.. Yang. R. (2013). Martinez-Botas. For steady state conditions.. M. D. While the energy averaged values match reasonably. Ehrhard.9 CONCLUSION AND OUTLOOK A novel and well validated direct shaft torque measurement technique has been employed for turbine stage performance assessment under steady state as well as under engine conditions. USA.. Texas. In a step-by-step process. B. the achieved experimental results were compared with CFD data. Texas. GT2013- 94538 [2] Lüddecke. The results can be employed to validate or (in case of significant deviations) improve state of the art turbine map extrapolation methods and strategies. ASME Turbo Expo 2013: Power for Land. “Contactless Shaft Torque Detection For Wide Range Performance Measurement of Exhaust Gas Turbocharger Turbines”.. Steinacher. Sea and Air. if these are “rebuilt” virtually within an engine simulation environment. Proc. San Antonio. Bargende.. B. (2012). San Antonio. Furthermore. USA. Proc. The obtained results represent a fundament for the deeper understanding of turbocharger turbine behaviour not only under engine-like or pulse flow conditions. Under pulsed flow conditions. Copeland. the authors are grateful for the excellent support of Mr Frederik Haußmann and Mr Marco Leonetti from the FKFS in Stuttgart. Texas. J. “On wide mapping of a mixed flow turbine with regard to compressor heat flows during turbocharger testing”. P.. M. B.. based on simple correlations for average gas flow velocity and the speed of sound.. the obtained unsteady results initially differed significantly from the measured steady state values from hot gas stand... Finally. The data can be utilized to judge the quality of engine simulation results of the same operating points. (2013) “Radial Turbine Rotor Response to Pulsating Inlet Flows”. but under real engine conditions. From the results presented it can be concluded that the aerodynamic behaviour of the investigated turbine stage can be assumed to be quasi-steady. GT2013-95182 [4] Newton. Romagnoli. Bargende.. ACKNOWLEDGMENTS The authors would like to thank Mr Philipp Nitschke from ICSI in Heidelberg for his valuable contribution to this work. Seene. Xu. Filsinger. D. especially the characteristics in terms of the slope of the curves and the expected pressure ratio for zero torque showed deviations. exhibiting a good qualitative as well as quantitative agreement.. Unsteady effects seem to be rather small and therefore are negligible. Proc. B.[5] Chiong. (2011) “A Revision of Quasi Steady Modelling of Turbocharger Turbines in the Simulation of Pulse Charged Engines”. Proc. Ehrhard. R. D. H. P. J. GT2006-90348 [12] Capobianco. (2010) “Extension of performance maps of radial turbocharger turbines using pulsating hot gas flow”. ASME Turbo Expo 2006: Power for Land. Lüddecke. W. Proc.. M. J. Ninkovic. Sea and Air... W.. J..... J.. N.. Texas. 17th Supercharging Conference. Lüddecke. P.... A.. H.. Proc. “Der neue 2. T. Hoffmann. (2012).. Proc.. Rückauf.. Filsinger.Hatz (2008). Aachener Kolloquium Fahrzeug. Proc. Kaufmann. R. A... Elsäßer. (2009). S. Bargende. W. [13] Reuter. F.. Sea and Air...und Motorentechnik [8] Wieske. Taylor. Hoffmann....enginos. Uhlmann. (2009). Thermodynamic Package for Internal Combustion Engines. Enginos GmbH. Scharf. ASME Turbo Expo 2013: Power for Land. Vedder... (2010) “Experimental investigation into the pulsating flow performance of a turbocharger turbine in the closed and open waste-gate region”. Martinez-Botas. B. “New Concepts for Optimising Transient Torque Response and Fuel Economy of Turbocharged Gasoline Engines”. M. S. Rajoo.. Costall. R. Bin Wan Salim.. Spain. (2013) “Assessment of Cycle Averaged Turbocharger Maps through One Dimensional and Mean-Line Coupled Codes”.de [11] Costall. R.. B. A. A. 9th International Conference on Turbochargers and Turbocharging of the IMechE. Barcelona.. Marelli. San Antonio. S. 18th Aachen Colloquium [9] Wieske. GT2013-95906 [6] Lüddecke. Dresden 320 . Szymko.-I.0l TFSI mit Audi Valvelift System für den Audi A4 – die nächste Generation der Audi Turbo FSI Technologie”. M.. A. 17. Fraser. F. Böhme. MTZ worldwide Edition: 2009-11 [10] Tiger Version 2. Wedowski. Elsäßer. Ewert.. Proc.. 16th Supercharging Conference. R. A. Martinez-Botas. A. S. Dornhöfer. Filsinger. S. Proc. S. S.. D. J. Budack. 9th International Conference on Turbochargers and Turbocharging of the IMechE. “Heat transfer correction and torque measurement for wide range performance measurement of exhaust gas turbocharger turbines”.. R. www. Dresden [7] Wurms. (2006) “Assessment of Unsteady Behavior in Turbocharger Turbines”. S. Romagnoli. Eiser. USA. D. S. R. A. Ewert. Koch. “Optimisation of Gasoline Engine Performance and Fuel Consumption through Combination of Technologies”. [14] Aymanns. Unequal admission NOMENCLATURE Π Pressure ratio ߙ Absolute angle (°) ݉ሶ Mass flow (kg/s) ߚ Relative angle (°) ܣ Cross sectional area (mm2) ௩ Vane throat distance (mm) ܴ Radius to area centre (mm) ݒ݊ݏ Vane pitch (mm) ܿ Absolute flow velocity (m/s) ݈݊ݒ Vane chord (mm) ߩ Density (kg/m3) Efficiency ߠ Circumferential angle (°) ܾ Inlet width (mm) ߰ Azimuth angle (°) VGT Variable Geometry Turbocharger EGR Exhaust Gas Recirculation SPR Scroll Pressure Ratio CFD Computational Fluid Dynamics NOx Nitro oxygen Subscript 0 Total/Stagnation condition L Large scroll 1 Volute inlet S Small scroll 4 Vane exit D Double-entry ݎ Radial direction ts Total-to-static ߠ Circumferential direction _______________________________________ © The author(s) and/or their employer(s). 2014 321 .Performance and flow-field assessment of an EGR pulse optimised asymmetric double-entry turbocharger turbine M Sakai. an asymmetrically divided double-entry turbine was developed to respond to the imbalance of mass flow due to EGR extraction from one side. Pulse turbocharging. R F Martinez-Botas Department of Mechanical Engineering. UK ABSTRACT In this paper it will be presented a novel turbine concept specifically designed for exhausts pulse flow energy conservation and EGR control. The EGR side was equipped with variable geometry vanes in order to control the EGR rate and to optimize the flow entering the turbine wheel. A detailed analysis on the design and efficiency of the asymmetric turbine is provided in this paper. Imperial College London. In order to combine both these features. Computational Fluid Dynamics. Variable Geometry Turbocharger. whereas no vanes were contemplated in the other turbine entry. Keywords: Double-entry turbine. A Romagnoli. Exhaust Gas Recirculation. single-entry VGTs are normally used due to their ability to support EGR and instantaneous boost pressure response.e. VGT also offers the additional advantage of varying the inlet area to the turbine wheel thus maximizing exhausts flow energy extraction at different engine operating conditions. Inner Outer limb limb Entry 2 Entry 1 (a) (b) Figure 1: Comparison between multiple-entry turbine designs (1): (a) Twin-entry and (b) Double-entry turbine design In order to merge the requirements of imbalanced mass flow and the pulse turbocharging advantage. However. double-entry turbine feeding area is divided circumferentially and incoming flows are guided with radially doubled scrolls. In order to extract exhaust gases and recirculate it into the intake side. In order to avoid exhausts pulse interference. aim of this research. (2) and lately optimized by Brinkert et al. Therefore a VGT is necessary to increase exhaust manifold pressure corresponding with widely changeable transient operation. these are isolated with multiple-entry turbines. turbocharger is inherently one of the most promising enabling technologies towards achieving engine design for low emissions and fuel consumption. Pulse turbocharging (i. In large multi-cylinder engines. conserving pulse energy until the rotor entry) is a well- known technique to utilise the maximum energy of the exhausts. The reasons for the selection of this turbine configuration are explained as follows. is that of combining the advantages of exhausts pulse energy extraction and variable geometry turbocharging in one single turbine design. exhaust manifold pressure has to be higher than the intake manifold pressure. The choice of asymmetric 322 . with the design of an asymmetric variable geometry double-entry turbine. even though the average exhausts back-pressure is lower than the charge air pressure thus limiting the operations for this design.1 INTRODUCTION AND PROBLEM DEFINITION Due to the increasingly stringent emission regulations and demand for high fuel economy. Turbochargers are not only expected to provide efficient exhausts energy recovery. Besides supporting EGR control strategies by changing exhaust manifold pressure. The twin-entry turbine is divided meridionally and each incoming flow is fed into the entire rotor circumference. asymmetric vaneless twin-entry turbine design has been proposed by Müller et al. A different approach than that provided by (2) and (3) is proposed in the current paper. Two types of multiple-entry turbines are currently available in the market: meridionally divided “twin-entry” turbines and circumferentially divided “double-entry” turbines (Figure 1). Their work showed that it is possible to achieve remarkable EGR-rates in some regions of the engine map. In contrast. the disadvantage of single- entry VGTs is that they do not enable to maximize energy extraction out of the exhausts pulses (pulse turbocharging) since there is no pulse separation within the turbine volute (as in multiple-entry turbines). but they are also used in order to support engine operation by controlling engine back-pressure for enhanced EGR rate. (3). Hence. the performance of the large vaneless scroll was calculated as conventional fixed geometry turbine. This flow imbalance can hardly be controlled in twin-entry turbines since the flow leaving the two entries mix together before entering the wheel. This is not the case in double-entry turbine configuration since the flows from the two entries are completely isolated and introduced into the turbine wheel separately. The two turbine inlets were treated as two independent turbine scrolls. A typical arrangement for the asymmetric variable geometry double-entry turbine is given in Figure 2. More details are provided in the next paragraph. the efficiency of the small and large scroll turbines were individually calculated using a mean-line model developed at Imperial College (4). Variable geometry vanes were only contemplated at the end of the scroll in which exhausts are extracted (small scroll in Figure 2) whereas in the other scroll no vanes were included. the presence of variable geometry vanes is made necessary in order to support EGR operations. the small vaned scroll was treated as single-entry VGT. less losses). vane pressure loss and vane clearance leakage loss) but still supporting EGR operations. the inlet flow angle) in order to maximize the exhausts flow energy extraction. As the EGR rate varies (hence varies) the variable geometry vanes will optimize the inlet flow area to the turbine wheel (i. In order to calculate the asymmetric double-entry turbine performance.double-entry geometry in place of a twin-entry one can be explained by considering that adopting symmetrically divided multiple-entry turbines in EGR engines.e. The overall efficiency of the asymmetric double-entry turbine was then calculated as mass flow weighted average of the efficiencies of each turbine scroll. In the small scroll instead. + = (1) + 323 . In contrast. With the asymmetric variable geometry double-entry turbine housing. so that the flow controllability is believed to be more effective than in twin-entry turbine.g. the turbine wheel will mainly be driven by the large scroll flow which is optimised for higher efficiency over wide turbine operating conditions (no vanes. Since the impact of EGR in the large scroll is negligible. EGR Vaneless valve Large scroll EGR cooler Vaned Small scroll Figure 2: The system of the Asymmetric double-entry turbine with a 6-cylinders engine 2 PRELIMINARY ANALYSIS The asymmetric double-entry turbine design started with preliminary off-design performance analysis. would lead to an imbalance of mass flow caused by EGR extraction from one side of the exhausts manifold. The main purpose of this partial vane arrangement is to minimise the losses due to the presence of vanes (e. 1 0.4 0. the efficiency of several configurations were assessed and compared with that of an equivalent single-entry VGT with same turbine wheel. ߟௌ and ߟ are individually calculated turbine efficiency of small and large scroll.2 Conventional VGT Single-entry VGT Asymmetric double-entry Double-entry 160:200 turbine 1 Mass weighted efficiency t-s 30% 20% 10% EGR 40% 0.8 0.7 0.3 0.6 EGR 0% 0. an EGR rate of 10% was chosen as initial design value (refer to the text box in Figure 3).9 1 1. 3 ASYMMETRIC DOUBLE-ENTRY DESIGN The design of the asymmetric variable geometry turbine scroll started with the selection of the EGR rate and the circumferential division to be considered at the design-point (10% and 160:200 respectively as described in the previous paragraph). the design procedure revolved around the assumption that the incidence angle at the end of small scroll should be identical to that at the end of large scroll. 324 . an improvement of 3% in efficiency could be found for a wide range of EGR rates (from 0% to 40%). 1. The mean-line method results showed that it is possible to achieve better turbine efficiency than conventional VGTs when EGR is being operated widely.2 Asymmetric double-entry preliminary design features Circumferential division: 160:200 .2 0. 1 In an asymmetric double-entry turbine.where.EGR 10% 0 0 0.8 3% 0. After a preliminary analysis run on a typical duty cycle of an off-road engine.6 0. In order to choose the optimum circumferential division1. Once these two parameters were fixed. This can be achieved by varying the vanes angle in the small scroll in order to match the best flow angle at the inlet to the turbine wheel and hence maximize exhausts flow energy extraction. The analysis showed that the optimum circumferential division varies depending on different EGR rates and the final choice fell on a 160:200 arrangement (refer to Figure 4).1 1. there are same radii and widths of the flow path but different two circumferential flow areas for small and large scrolls (Figure 4).5 0.4 0.2 Normalized Inlet Mass flow to the Turbine Figure 3: The turbine efficiency advantage of asymmetric double-entry turbine against conventional VGT when EGR is being operated widely (The mass flow was normalized with the design-point mass flow rate of the single-entry VGT used for comparison) This is shown in Figure 3 where for the selected turbine division. From the free vortex correlation. = ⁄ · = ⁄ · (7) ⁄ = (8) ⁄ From equation (8). the ⁄ values were set to obey to the free vortex condition following the single-entry VGT designed at Imperial College. · · = (5) · · For incompressible flow. In order to perform a comparison with a variable geometry single-entry turbine. it can be gathered that for a set EGR rate of 10%. a turbine wheel designed and tested at Imperial College by Abidat (5) (and lately used by Rajoo (6) for single-entry VGT design) was chosen. ∙ ∙ = ∙ ∙ (3) By substituting equation (2) and (3) into the following absolute turbine inlet angle. the ⁄ ratio between the large and small scrolls can be expressed as in equation (9). ⁄ % = = (9) ⁄ % Then it is now possible to separate circumference into two. ( + = 2 ). the relationship between the centroid radius of the area and flow velocity is: ∙ = ∙ (2) The continuity equation for incompressible flow provides the relationship between radial flow velocities at the inlet and exit to the volute (station 1 and 4 respectively). = ⁄ · (6) In case of single-entry turbine. whereas in a double-entry design turbine the inlet angles should be considered separately and identical to each other. In order to run a one- to-one comparison. The key features for the VGT asymmetric double-entry turbine are shown in Table 1 and Table 2. the equation (5) can be simplified. In other words. and set the ratio between the two ⁄ s identical to circumferential division. = (4) We obtain. 325 . despite the asymmetric double-entry turbine volute is divided into two. the sum of the small and large volute ⁄ values and their ratio against azimuth angle (Figure 4) were designed to be identical to that of the single-entry VGT designed by Rajoo (6). = 2 since all the incoming flow is distributed around the circumference of the turbine wheel. 33 mm Large scroll 16. Table 1: Design conditions Parameter Pressure ratio 2.07 mm (reference diameter) Number of blades 12 Number of vanes 9 Vanes angle (standard vanes angle) 67.65 ° Vane pitch angle 20 ° 326 .678 kg/s Rotational speed 60000 rpm Target EGR 10 % 40 Small scroll Large scroll 35 30 A1/R1 [mm] 25 20 15 10 5 0 0 40 80 120 160 200 240 280 320 360 Azimuth Angle [deg] Asymmetric double-entry turbine Single-entry turbine by Rajoo 2007 (a) Station 4: vanes exit (downstream) Small scroll Station 3: vanes exit Stationedge) (trailing 3 Station 2: vanes inlet 160 200 Large scroll Station 1: volute inlet (b) (c) Figure 4: Asymmetric double-entry turbine volutes A/R design (a) Comparison of A/R change along the volutes from its tongue between asymmetric double-entry turbine and single-entry turbine (b) Asymmetirc double-entry turbine layout (c) Single-entry turbine layout for comparison Table 2: Geometric details of asymmetric double-entry turbine Geometric feature Asymmetric circumferential division 160 : 200 A/R Small scroll 13.91 Inlet temperature 344 K Mass flow rate 0.67 mm Radius Tongues 70 mm Turbine wheel 42. 577.75 ~ 0.85.045 8.Unlike single-entry VGTs in which the vanes are arranged around the entire turbine wheel circumference. = ⁄ (10) whereas the latter is a function of the optimum tangential lift coefficient which in the Zweifel’s criterion (7) is suggested to fall between 0. Wedges) Structured (Hexahedral) and Vane section 78. = | − | (11) Unlike the single-entry turbine housing designed by Rajoo (6).868 (Tetrahedra.416 1.956 (Hexahedral) Unstructured Exit ducting 39. In addition to this. The former is given as function of vane geometries (8). 4 CFD ANALYSIS 4. here defined Scroll Pressure Ratio (SPR).684 Unstructured (Tetrahedra.121 (Tetrahedra.180 75.e. 20° and 10°).246. In the 160:200 circumferential division.372. the vane pitch is also dependent on the exit flow angle and the Zweifel’s criterion (7).1 Computational analysis and discussion In order to understand the turbine basic behaviour of the asymmetric double-entry turbine. the angle between two vanes is required to be a common divisor of the asymmetric circumferential division angles in order to have identical distance from tongue to vane. has been introduced (equation 12) to provide the rate of imbalance between the small and large scroll. in a double-entry turbine configuration. Wedges) Structured Rotor domain 1. Wedges) Total 1. 40°.323 120. the simulation results have been obtained with CFD analysis. the number of vanes is strictly related to the location of the two tongues.629 The turbine performance analysis was run under equal (same pressure ratio within the turbine inlets) and unequal admission (unbalance of pressure ratio between the inlets) conditions. If the vanes are equally spaced around the periphery of the turbine wheel. For ease of discussion a parameter. = = (12) 327 . the vanes shape was changed from straight to curve in order to maintain the vane angle from the leading edge to the trailing edge and therefore obtain the same flow angle (in station 4) as that in the vaneless section. the angle between two vanes should be a common divisor between 200° and 160° (i.938.964 1. Table 3: Mesh characteristics of turbine domain Number of Number of Region Element type Elements Nodes Unstructured Volute 575. The CFD analysis was conducted using commercial software ANSYS-CFX and in Table 3 it is provided the mesh characteristics of the whole turbine domain. 0 3.0 2.0 4.2 1.6 0.5 or 1.0 1.0 3.8 0. independent of which limb is flowing less mass flow (either SPR 0.2 1.2 0.4 0.0 2.0 1.3 0.2 0.5 0.8 0.0 2.3 pressure ratio for 50% speed and 78% at 2.1 (b) 0.2 0. (b) 100% speed and standard vane position (The mass flow was normalized with the design-point mass flow rate of the single-entry VGT used for comparison) The analysis started by setting the vane angle at design-point (refer to Table 2). the turbine performance decreases significantly.5 0.6 0.0 1.0 3.0 1.0 Pressure Ratio t-s Pressure Ratio t-s 1.8 Turbine Efficiency t-s Turbine Efficiency t-s 0.0 0.5 present large efficiency drop).8 0.0 3.3 0.4 0.4 0. in which it is believed large amount of EGR is likely to be required). This is consistent with the initial design assumption of peak efficiency point occurring at 2.43 pressure ratio at 100% speed.6 0. 328 .0 4.4 0.6 0.1 (a) 0.2 (a) (b) 0.0 4.91 pressure ratio (5) (6) (refer to Table 1) and also with previous available literature showing that in multiple-entry turbines the peak efficiency point occurs under equal admission (9) (10) (11).7 0. 1.0 1.0 2. The simulation results are given in Figure 5 and show that the highest turbine efficiency occurs under equal admission conditions with an efficiency value of 76% at 1.0 0.9 0.0 30000rpm (50% speed) 60000rpm (100% speed) 0.0 Pressure Ratio t-s Pressure Ratio t-s Figure 5: Turbine performance of the Asymmetric double entry turbine at (a) 50% speed and standard vane position.0 4. considering tow rotational speeds of 60000rpm (100% design-speed) and 30000rpm (50% design-speed.0 Normalized Mass Flow rate Normalized Mass Flow rate 0.0 1. As the rate of imbalance between the two inlets increases.9 0. and five different values for the SPR.7 0. 0 0. the mass flow in the small scroll would experience a blocked flow condition (dashed line. flow type C) and in an extreme case some backflow could occur (flow type D). 4. As the pressure in the small scroll keeps decreasing as consequence of larger and larger EGR rates. This is not the case in a real engine since the exhausts flow are instantaneously pulsating. typical mass flow patterns are illustrated in Figure 6 for 50% turbine speed. ○ B: . Hence the flow type B (blue shaded area in Figure 6) cannot be obtained in the steady-state condition of asymmetric double-entry turbine design.2 SPR 1. 329 .5 2. ○ D : 0 In order to understand the limit of mass flows through which the asymmetric double-entry turbine can operate. Hence flow type B. Starting with steady-state flow assumption. trying to include also the effects of pulsating exhausts flow conditions. the mass flow in large scroll can be exceeded by that in the small scroll as shown from the black marks in Figure 6. due to the firing order of the engine.5 2.5 3.0 SPR 0.8 PRsmall SPR 1.0 SPR Figure 6: Small turbine volute flow interaction: ○ A : Steady-state operation.5 SPR 1. Therefore over an entire engine cycle.0 2. the mass flow in the small scroll can never exceed that in the large scroll due to the cylinder distribution (3 cylinders connected to the small volute and EGR circuit. and other 3 cylinders connected to the large scroll: ).5 Blocked flow in small scroll small scroll mass flow Blocked flow in large scroll cannot exceed large scroll mass flow SPR 0. it is worth noting that the results of Figure 6 are true only under the assumption of steady-state flow.0 1. However.○ C: . from A to B.0 B C A 1. However this will not be discussed further within this paper which is currently looking at steady- state efficiency assessment for the asymmetric double-entry turbine design. it can be noticed that in contrast with standard flow type A.5 1.0 30000rpm (50% speed) available under pulsating flow condition only 3. In the large scroll instead the mass flow can never be exceeded by that in the small scroll since in the large scroll there is not EGR flow extraction.5 C Backflow in Backflow in small scroll D D large scroll 1. In the figure have been identified four flow areas.0 0. flow type C and flow type D in the large scroll can only be obtained in an experimental lab set-up but it would not occur during standard engine operating conditions since there is no EGR flow extraction. 0 0% 10% 20% 30% 40% 50% 60% 70% 80% EGR rate Figure 7: Comparison of optimum EGR rate: (a) at 50% speed and -10° of vanes angle (b) at 50% speed and standard vanes angle (design-point) (c) at 50% speed and +10° of vanes angle In Figure 7 it is reported a comparison between efficiencies at 50% speed for different vanes angle position.2 Turbine performance and optimisation with vanes angle From the previous analysis it was found that the highest efficiency point occurs under equal admission conditions. EGR rates and SPR values. If the vanes angles were not optimized to achieve equal admission condition. the optimum EGR rate can be adjusted by varying the vanes angle position.2 (a) 0. 1.8 0.2 (c) 0.0 Turbine Efficiency t-s 0.4.4 0. 330 .0 0% 10% 20% 30% 40% 50% 60% 70% 80% EGR rate 1. the turbine efficiency would drop due to the imbalance of mass flow between the two inlets.0 0% 10% 20% 30% 40% 50% 60% 70% 80% EGR rate 1.6 0.6 0.8 0. EGR rate is expressed as in equation 13.4 0. In other words.8 0.4 0.0 Turbine Efficiency t-s 0.0 Turbine Efficiency t-s 0. By changing the vanes angle when target EGR rate is changed.6 0. it is possible to keep higher turbine efficiency.2 (b) 0. Then the complete turbine design moved into 3-D modelling and CFD performance assessment. as shown in Figures 7a and 7c. The benefit of variable vane configuration could be appreciated for different EGR rates where the possibility to optimize the flow condition at the inlet to the turbine wheel showed that it is possible to retain an equal admission conditions (and hence optimum turbine efficiency) for a wide range of EGR rates. Simulations were run for a number of different turbine speeds (100% and 50% design-speeds). At design-point vane position (Figure 7b). is totally depending on the variable geometry vanes angle in Figure 7. As the SPR decreases slightly (SPR 0. and therefore in order to maintain optimum turbine flow conditions (corresponding to equal admission conditions). As final remark about this project.5 to 1. with peak efficiency value of 76%. 5 CONCLUSION In this paper. The simulation results showed that the peak efficiency point occurs under equal admission (SPR 1. In some engine operating conditions. large EGR rate is not required. As a result of this process.5) and vanes angles (-10°. It is obvious that the equal admission condition line. the vanes angle need to be varied. the final turbine housing arrangement comes as asymmetric double-entry with circumferential division of 160:200 for 10% EGR rate (design-point) and variable geometry vanes only in the turbine inlet where EGR flow extraction occurs. 2 Despite 10% target EGR rate the optimum EGR rate calculated by CFD was 20%. This may require some further investigation and analysis for the vanes design and profile. The design started with preliminary mean-line analysis in order to fix the basic turbine geometrical parameters. Thanks to vanes position adjustment no much penalty in efficiency was observed at 50% speed.8) in order to maintain the same level of efficiency within the turbine. ±0°. it is notable that the equal admission condition is not always available since the intake manifold pressure and the exhaust manifold pressure are always changing on engine operation map. the design and performance assessment of a novel asymmetric double-entry variable geometry turbine was discussed. The test programme will focus on the validation of the presented CFD results as well as in pulsating flow performance assessment. SPRs (from 0. the optimum EGR rate is approximately 20%2 as shown by the black bold line.0) with 78% turbine efficiency at 100% design speed and 20% EGR rate. an equal admission condition can be obtained for no excessive EGR rate values (10% and 30% for -10° and +10° vane position respectively). − = = (13) + + × Two additional vanes position (±10° of the design-point vanes angle) were considered in the analysis. the EGR rate should increase significantly (more than 30% EGR). which gets always maximum efficiency. By slightly varying the vanes position (±10° of the design-point vanes angle). This novel turbine housing design was conceived for improving engine EGR operation and enhancing exhausts pulsating flow energy extraction in large multi-cylinder engines. However. 331 . it is worth saying that a prototype of the asymmetric variable geometry double-entry turbine will soon be tested at Imperial College London. and +10°). Martinez-Botas. Proceedings of Institution of Mechanical Engineers. 53. N. Proceedings of ASME Turbo Expo 2011. pp. Müller. Ph. Martinez-Botas. and Baines. and Seiler. D. “Design and Testing of a Highly Loaded Mixed Flow Turbine”. “Understanding the Twin Scroll Turbine-Flow Similarity”. “Performance Prediction of a Nozzled and Nozzleless Mixed-Flow Turbine in Steady Conditions”. 2011. 10. 178. 8. G. A. 2012. Thesis. 2008. GT2008-50614.H.. F. Abidat. et al. 60. Thesis. 557-574. 2007. M. 7... “Steady and Pulsating Performance of a Variable Geometry Mixed Flow Turbocharger Turbine”. R. I.. 11. “Experiments Concerning the Aerodynamic Performance of Inward Flow Radial Turbine”. 2011... 332 . ASME Journal of Turbomachinery. Japikse. Technology and Medicine. Hiett. 4. “Introduction to Turbomachinery”. Proceedings of ASME Turbo Expo 2008. D. “Turbocharging the Internal Combustion Engine”. 1984. Pt 31(II). R. and Johnston.D. “Comparison Between Steady and Unsteady Double-Entry Turbine Performance Using the Quasi-Steady Assumption”.. and Seiler.D. International Journal of Mechanical Sciences. “Unsteady Performance of a Double Entry Turbocharger Turbine with a Comparison to Steady Flow Conditions”. M. 2011. 133.REFERENCE LIST 1.. Copeland. Proceedings of ASME Turbo Expo 2008. Technology and Medicine. and Martinez-Botas. 3. pp. ASME Journal of Turbomachinery. A. et al. R. 6. Romagnoli. Romagnoli. Watson. 5. 2. F. 031001.. 1982. Vol. et al... M. 1963. D. Ph. and Janota M. 1991. “Comparison between the Steady Performance of Double-Entry and Twin-Entry Turbocharger Turbines”. Imperial College of Science. C. Copeland. GT2011-46820. pp. 9. N. “The Asymmetric Twin Scroll Turbine for Exhaust Gas”. 2011. F. Brinkert. N. F.. C. S. Vol. S. GT2008- 50827. Imperial College of Science. Rajoo. 6-17. M. The results show that the turbine with nozzled volute has higher efficiency at low load conditions in both steady and unsteady conditions.Comparison of the influence of unsteadiness between nozzled and nozzleless mixed flow turbocharger turbine M Y Yang1. NOMENCLATURE IC Internal combustion Cr Radial component of velocity m/s Cϑ Tangential component of velocity m/s C velocity m/s U Rotor tip velocity m/s PR Pressure ratio P pressure Pa MFP Mass flow rate parameter kg/s. UK 2 Universiti Teknologi Malaysia. reflecting the filling and emptying as well as the wave action effect in the turbine. China ABSTRACT Nozzled and nozzleless volutes are the most commonly used stator configurations for the turbocharger turbine. Malaysia 3 State Key Laboratory of Automotive Safety and Energy. The improvement of the cycle average efficiency for the nozzled volute under pulsating conditions is influenced on the pulse frequency. Instead. the results of the 1-D model show that the swallowing capacity curve has an evident influence on the unsteadiness in the turbine. W L Zhuge1. Tsinghua University. The steady peak efficiency is nearly the same as the nozzleless one but it shifts to higher velocity ratio. Furthermore. the unsteadiness of a turbine is enhanced by reducing the swallowing capacity or increasing the turbine loading. are similar for two configurations.2. The hysteresis loops of the swallowing capacity. S Rajoo2 1 Imperial College London. M H Padzillah1. The comparison shows that the volute configuration has no direct influence on the ‘unsteady’ behaviour in the turbine. 2014 333 . In this paper detailed experimental investigation was carried out to compare the performance of a mixed flow turbine with nozzled and nozzleless stator under steady and pulsating conditions which replicates the pulsating exhaust flow from the internal combustion engine. √ /Pa m Mass flow rate kg/s _______________________________________ © The author(s) and/or their employer(s).3. R F Martinez Botas1. Generally there are two types of the turbine identified by the configurations of the volute. 334 . including nozzleless turbine and nozzled turbine. The swallowing capacity is remained similar for two configurations by adjusting the vane geometries. The efficiency of the nozzled one drops gradually below its nozzleless counterpart as the operation point deviates from the peak point. 3]. 2. The volute with a removable nozzle was used on a mix flow rotor in the testing. It was implied that the different observation from conventional conclusions was attributed mainly to different surface finishing when the comparison was made. The different flow pattern at exit of the volute was considered to be the main reason for discrepancy of the turbine efficiency [1. On the other hand. The turbine with nozzled volute was claimed in most of the researches to have better peak efficiency than the one with nozzleless stator. thus produces a more uniform flow distribution at the rotor inlet. The nozzleless turbine is usually applied for the turbocharger of the IC engine in passenger vehicle due to its simplicity (cost) and reliability. A direct performance comparison between the two volute configurations in a radial turbine was conducted via experimental and CFD method [4. in relatively larger IC engine (10-15L). CFD results showed that the flow was less uniform in the nozzled volute due to wake and jet caused by nozzle vanes. the experimental comparison was also carried out on a mix flow turbine [6].m Subscripts 0 ambient parameter ave cycle average in Inlet out Outlet V nozzled (with vane) VL nozzleless (vaneless) INTRODUCTION The increasing demand of vehicle with low CO2 emission has forced the industry towards highly efficient engines. η efficiency β relative flow angle deg. While the flow distortion caused by the volute tongue is damped by the nozzle. A turbocharger turbine plays a key role in recovering the energy from the exhausted gas of an engine and hence attracts intense researches targeting at the performance improvement. torque N. 5]. Except for the investigations on the radial turbine.Greek letters α absolute flow angle deg. and higher pressure loss was produced in the nozzled volute due to the introduction of vanes. Baines and Lavy [1] indicated that nozzled turbine offers superior efficiency compared to nozzleless turbine near the peak efficiency point. and hence results a turbine with smaller size. The experimental investigation of the flow pattern at the exit of volute showed that highly distorted flow appeared in the nozzleless volute especially at the region near the tongue. The efficiency of the nozzled turbine shows constantly better efficiency than the nozzleless one. The performance comparison under steady conditions between the turbine with nozzled and nozzleless volute has been proceeded extensively by experimental or CFD method in literatures. The results showed that the nozzleless volute always has superior efficiency at all operating conditions. a nozzle ring is usually employed to turn the flow in limited space in the turbine stator. Rajoo [9] conducted detailed experimental comparison of unsteady performance of a nozzled turbine and the results demonstrated that both the shape and the area covered by the hysteresis loops varied with the nozzle vane angle. Furthermore. Layout of the test rig. of which the maximum output power and speed can be achieved at 60 kW and 68 kRPM respectively. the vanes are removed and strut holes are sealed by the same base of vane. The nozzled and nozzleless volute configurations for comparison are shown in figure 2. The nozzle was considered to shield the rotor from the pulsating flow and produced a different hysteresis loop when the opening of the nozzle changed. Both vane angle and number of the nozzle are adjustable. The pulsated air flow is produced by the pulse generator which produces a similar pulse shape as the engine exhausted gas. The main geometries of each component are listed in table 1. The layout of the rig and the turbine for investigations are shown in figure 1. For the nozzless turbine testing. The output power of turbine is absorbed by electrical eddy current dynamometer. The current research experimentally investigates the influence of the volute configuration on a mix flow turbine performance under steady and pulsating flow conditions. Figure 1. 10]. volute and the mix flow rotor The investigated mixed flow turbine with the leaned vane nozzle was developed by Imperial College [8. The cycle averaged efficiency of the nozzleless volute superiors the nozzled one. 335 . as shown in sub-figure (b). [8]). The turbine is driven by the compressed air and heated to 65 degree centigrade by an electrical heater. Romagnoli [7] firstly experimentally compared the unsteady performance of a turbine with nozzleless and nozzled volutes.Few investigations have been conducted on the performance comparison under pulsating conditions between the nozzled and nozzleless turbine. 1 TEST FACILITIES AND EXPERIMENTAL METHOD The test was carried out on the rig in Imperial College London. It is noteworthy that the swallowing capacity of the two configurations was not matched for the comparison and the turbine with nozzled volute was not working at the optimized vane angle (60 degrees. The results showed that the hysteresis loops of the swallowing capacity were quite different between two configurations and the vanes of the nozzle were considered to be the reason for the discrepancies. the influence of the swallowing capacity on the unsteadiness in the turbine is demonstrated via the well calibrated 1-D unsteady method. 336 . For the steady test. The instant static pressure are measured at the measurement plane and turbine exit by Schaevitz pressure transducers P700 and instant temperature are evaluated by T-type thermocouples together with the instant pressure. Main geometries of the mix flow turbine Stator A/R of the volute 33 mm Vane number of nozzle 15 Rotor Mean inlet radius of rotor 41. The instant mass flow rate is measured by Dantec 1-D hotwire 55P16 and the device is calibrated by the V-Cone flow meter under steady conditions over the range of the mass flow and temperature covered by a pulse.5 mm Tip radius at rotor exit 39.8 mm Inlet blade angle at middle of blade 20 deg. middle and high loads) at the two speed lines. The mass flow rate is measured by McCrometer V-Cone Flow meter (dP meter) of which the accuracy is ±0. The output power is evaluated by the torque from dynamometer as well as the rotor rotational speed. Then the swallowing capacity of the nozzled turbine is adjusted to match with the baseline via changing the vane angle. Cone angle of blade 50 deg. Configurations of nozzled/nozzleless volute The comparison between the configurations with nozzle and nozzleless is carried out under the same swallowing capacity characteristic. 40HZ and 60HZ) and three different loads conditions (low.5%. In order to achieve the similar swallowing capacity. The same vane angle is set for the nozzled volute during the unsteady test. The pressure upstream the turbine inlet and downstream the turbine exit are evaluated by a scanivalve and the temperature is measured by T-type thermocouples about 600mm upstream the volute inlet (measurement plane). It is considered to be a fair comparison under this constraint since the inlet conditions imposed at the turbine inlet is the same for two configurations given the same velocity ratio or inlet pressure.3 mm Exit blade angle -52 deg. the performance of the turbine with nozzless volute will be tested as the baseline firstly. For unsteady test. The chopper plates are set to be fully open. (a) nozzled volute (b) nozzleless volute Figure 2. Instant torque was evaluated by the rotor acceleration plus the load cell reading. Hub Radius at rotor exit 13. the performance is measured at two speed lines as 30 kRPM and 48 kRPM. Table 1. the performance is measured under three different frequencies (20HZ. 731.2 COMPARISON OF THE STEADY PERFORMANCE Figure 3(a) shows the swallowing capacity of the two volute configurations at 30 kRPM and 48 kRPM. By inducing the nozzle vanes to the volute. the vane of the nozzle should be increased (αV > 337 . Steady performance comparison of two volute configurations Another important observation from figure 3(b) is the locations of the peak efficiency point. which is going to be demonstrated following. but the advantage of the efficiency for the nozzled one becomes obvious as the velocity ratio increases. the actual flow area at the rotor inlet will reduce due to the blockage of the vanes compared with the nozzleless volute. At 30 kRPM speed. as discussed below. As discussed before. The maximum difference of mass flow parameter is lower than 4% which exists at the highest pressure ratio at 48 kRPM turbine speed. This behaviour of the efficiency performance for the two configurations could be explained by the changes of the velocity triangle at the inlet of the rotor. The shifting can still be observed but less obvious for 48 kRPM speed partly due to the smaller measured range. respectively. the radial component of the velocity Cr has to be increased. (a) swallowing capacity (b) total-static efficiency Figure 3. the efficiency improvement by the nozzled volute is more evident as the speed increases from 30 kRPM to 48 kRPM: the peak efficiency is 0. the nozzled volute shows 5% efficiency improvement at 30 kRPM while 7% at 48 kRPM. In order to maintain similar swallowing capacity as the nozzleless configuration. thus the efficiency of the nozzled turbine can be expected to be good. the peak efficiency point moves from velocity ratio of about 0.55 to 0. The peak efficiency of the turbine with either volute is considered to be achieved at the same optimized incidence angle β at the rotor inlet.68 as the nozzled are installed.738 and 0. As the velocity ratio increases to the right side of the peak efficiency point. Under constraint of the same swallowing capacity (CrVL < CrV). and it is slightly higher for the nozzled volute than the nozzleless one at 48 kRPM. as shown in figure 4. Moreover. It can be seen that the swallowing capacity for the nozzled and nozzleless volutes at both speeds is matched properly over the whole measured pressure ratio range.69 for both configurations at 30 kRPM. taken the point at 0. The total-static efficiency for both volute arrangements at the two speeds is shown in figure 3(b). the matching procedure is conducted by adjusting the nozzle vane angle. The peak efficiency for two configurations is similar. The performance superiority of the nozzled volute over its nozzleless counterpart at the two rotational speeds is clearly demonstrated in the figure. The correspond vane angle of the nozzled volute is found to close to the optimized value (60 degrees).9 for instance. It can be seen that the employment of nozzle shifts the peak efficiency point towards high velocity ratio region. which is 0. the swallowing capacity manipulation in current research is achieved by adjusting the vane angle instead of the vane profile. It consequently results in a lower isentropic velocity for the nozzleless configuration thus the higher velocity ratio at peak efficiency point. especially at high velocity ratio conditions. at least for the investigated turbine. At off-peak conditions. respectively: = (2) 338 . Velocity triangle at rotor inlet for nozzled and nozzleless volute The efficiency comparison between the nozzled and nozzles volute shows that the conventional claim of the benefit at peak efficiency for the nozzled volute is not quite valid. the peak efficiency for two configurations is very similar. and hence the lower available power at the rotor inlet according to the Euler work equation. This could benefit the turbine performance in such a way that flow distortion caused by the volute tongue is damped significantly in the large interspace region. As a result. However. Different from the method utilized by Simpson [11]. On the other hand.αVL) to achieve the optimized incidence angle (β V = β VL). the peak efficiency achieved by two configurations is at similar level. 3 UNSTEADY TEST RESULTS 3.1 Cycle average performance The convenient way to compare the performance of different configurations under pulsating conditions is cycle average parameters via which the influence of the phase shifting methods can be avoided. the tangential component of the absolute velocity at the peak efficiency is lower for the nozzled volute (CϑV < CϑVL). As a result. It was usually considered that the uniform flow at the nozzle exit contributes to the higher peak than the nozzleless volute. the nozzleless volute in the experiment is produced by removing the flexible nozzle vanes from the nozzled volute. Figure 4. The shifting of the peak efficiency results from the vane angle adjustment and therefore the efficiency at low velocity ratio can be benefit from it. Instead. The cycle average efficiency is evaluated by the integration of actual power and isentropic power over a pulse cycle: ( ∙ ) = (1) ∙ ∙ ∙( ( ) The cycle average pressure ratio and mass flow rate are defined to evaluate the mean load of the turbine and the mean swallowing capacity under pulsating conditions. the pressure loss induced by the nozzle vanes reduces the turbine efficiency and therefore balances the benefit of the uniform flow. thus leaving large interspace region. the nozzled volute shows the efficiency advantage at most of the steady conditions. the loop expands dramatically indicating significant mass imbalance in the turbine. the nozzled turbine gradually shows the advantage over the nozzleless one. As the load reduces and the frequency increases. A B C (a) 30 kRPM (b) 48 kRPM Figure 5. The nozzled turbine has the biggest advantage at low frequency low load condition where the efficiency is 13% higher (location ‘C’). the advantage of the efficiency for the nozzled turbine is more obvious than at 30 kRPM. 339 . the loop shrinks as the frequency further increases to 60HZ. The loop follows the steady curve well except at high load conditions. the advantage of efficiency for the nozzled turbine under pulsating conditions is more evident at the higher speed. 3. = (3) Figure 5 shows the plots of cycle average efficiency at different loadings and frequencies at two rotational speeds. As the frequencies increases to 40HZ. which is the similar as the phenomenon at steady conditions. Moreover. At 30 kRPM the nozzleless turbine works more efficiently at high-load-low-frequency condition than the nozzled one (subfigure (a)). Specifically.1% higher at 60HZ low load condition (location ‘B’). but obviously the frequency imposes an evident influence on the magnitude. Figure 6 compares the loops at 48 kRPM middle load conditions for three frequencies at 20HZ. The shape of the loops is evidently influenced by the frequencies for both configurations: the loop at 20HZ is slim and stretches along in pressure ratio axis.2 Hysteresis Loops of the Swallowing Capacity The unsteady behaviour of the turbine under pulsating conditions is typically characterized by the hysteresis loops of its swallowing capacity.4% higher for the nozzleless one (location ‘A’).24) is 6. as shown in subfigure (b). As the speed increases to 48 kRPM. the efficiency of the nozzled turbine is 11. Figure 6-7 show the swallowing capacity loops of two configurations at different operation conditions. the efficiency at 20HZ high load condition (PRave=1. It is thus implied that the cycle average performance of the investigated turbines has similar behaviour with the steady performance. 40HZ and 60HZ. Cycle average efficiency at different frequencies/pressure ratios at 30/48 kRPM In generally it can be observed from both subfigures that the nozzled turbine is more efficient at low load conditions. It is noteworthy that the similar trend can be seen in the steady performance shown in figure 3(b). For instance. However. although the magnitude is obviously influenced by the frequency. The magnitude of the pulse also contributes to the unsteadiness in the turbine. It is demonstrated clearly that the load. quantified by the cycle averaged pressure ratio. it is clear that the variation of the loops at the corresponding load shows a very similar trend. which is reasonable to be evaluated by the area covered by the hysteresis loops and indicates the magnitude of mass imbalance during a pulse cycle. Comparing the loops between two volute configurations. The swallowing capacity loops of two configurations are compared for three different loads conditions at 40HZ. it can be seen that the shape as well as the ‘unsteadiness’ are very similar although 15 nozzle vanes has been installed in the nozzled volute. More importantly. It is inferred that the unsteady behaviour (filling and empty/ wave action) is not directly influenced to the volute configuration (nozzled/nozzleless). enhances the filling and emptying effect in the turbine. comparing the loops between two configurations. for both covered area and shapes. is not proportional to the frequency. Swallowing capacity loops at 48k RPM middle load different frequencies The ‘unsteadiness’ of the turbine. according to the research of literature [14]. (a) nozzleless (b) nozzled Figure 6. (a) nozzleless (b) nozzled Figure 7. Again the nozzle vanes have little influence on the unsteady behaviour in the turbine for those two configurations. Figure 7 shows the influence of the load on the unsteadiness in the turbine. Swallowing capacity loops at 48k RPM 40HZ different loads 340 . In general the loops extend outwards and move to higher pressure ratio along the steady curves as the load increases. One dimensional unsteady model for the turbine 341 . On the other hand. which is going to be discussed following. Ambient static pressure is used as outlet boundary condition. the ‘unsteadiness’ in the turbine under pulsating conditions is significantly influenced by the pulse frequency and turbine load. Instead.3 Influence of the swallowing capacity on the hysteresis loops In order to investigate the influence of the characteristic of swallowing capacity on the turbine unsteady behaviour. The turbine rotor is treated as a quasi-steady component and the swallowing capacity curves are employed to represent the rotor. This constraint of the swallowing capacity is also one of the key reasons for the phenomenon. in order to investigate the influence of the swallowing capacity on the unsteady behaviour. The configuration of different vane angle is modelled by the inlet/outlet area ratio of the nozzle. for which the reason can be partly explained by the research [12]. The geometries of the turbine system are simplified as one-dimensional pipes connected by the same topology. However. More importantly. The volute is simplified into three components in series: a nozzle to represent the part from inlet to the throat. 3. a nozzle to represent the nozzled or nozzles part at the volute exit. The configuration of the model is shown in figure 8. Test results including inlet total pressure and temperature are employed as the inlet boundary conditions. The length and inlet/outlet area (thus the volume) of the pipes are determined by the real geometries. a straight pipe to present the passage of the volute. it should be reminded that the swallowing capacity of the two configurations is the same during the testing. Figure 8. a 1-D unsteady non-viscosity model is developed for the investigation based on the in-house code ‘ONDAS’ [13]. the hysteresis loops are going to be compared at different swallowing capacity curves and different load conditions in this section. Two-step Lax-Wendroff scheme together with TVD method is applied to discretize the computational domains. Both the shape and the capsuled area of the loops are similar for the turbine with nozzled and nozzles volute as long as the turbine works at the similar unsteady conditions (in terms of pulse frequency and the mean load).Different from the conclusions drawn in the literatures. Therefore. the comparison discussed above demonstrates clearly that the unsteady behaviour of the turbine is not directly influenced by the volute configurations. the magnitude of the pulse is directly related with the unsteadiness of the turbine. the magnitude of the pulse is closely coupled with the frequency and turbine load in the experimental testing and can’t be decoupled. The shape thus the wave action in the turbine changes as well. thus the discrepancies of the loops are only caused by the turbine swallowing capacity. As the swallowing capacity drops (by closing the vane). representing the larger vane angle (close) and smaller vane angle (open). as shown in the figure 10(a). Nevertheless. Both the mass flow rate range and the shape of the loop are well predicted by the model. and the experimental data of different swallowing capacity (achieved by adjusting the nozzle angle) is not available. the loop expands significantly indicating enhancement of mass imbalance in the turbine. Validation of the 1D model at 40HZ middle load at 48 kRPM The predicted hysteresis loop is compared with the test results at 40HZ middle load at 48 KRPM in the figure 9. the turbine with smaller swallowing capacity is more ‘unsteady’ than the one with larger swallowing capacity. Figure 10 shows the hysteresis loops predicted by the unsteady model. but not as dramatically as the covered area. two curves of pressure ratio versus mass flow rate parameter are produced by mathematical extrapolation. thus this value is the same for all the vane angles. It is further inferred that for turbines confronted by a same pulse of the inlet pressure. the mass flow rate is well predicted in filling process but marginally under-predicted in emptying process. The influence of the swallowing capacity on the loops is clearly demonstrated in the figure. Figure 9. Using the point as the anchor. The volume of the pipes in the model and the approximation of the length of the volute are considered to be the main reasons for the discrepancy. resulting in a larger hysteresis loop compared with the test result. Specifically. The fillets and the blockage factor of the pipes are not modelled in the model which will result in a larger volume in the model. the 1D model is considered to be reliable enough for the investigation according to the prediction results. It should be noticed that the pulse magnitude and the pressure ratio range are the same due to the same inlet condition. 342 . As the main purpose of the discussion is to understand the effect of different steady flow swallowing capacity to unsteady turbine behaviour. Figure 10 (a) compares the hysteresis loops with different the swallowing capacity at the same inlet condition. By extrapolation the swallowing capacity curve is intersected with the horizontal axis and the value is the pressure ratio produced by the centrifugal force at the inlet of the rotor. the swallowing capacity curve of the original vane angle is increased by about 20% and reduced by 20% representing the open and close respectively. it is difficult to check the influence of the turbine load separately.As discussed in the test results. the turbine load has been shown to have significant influence on the unsteadiness. since the magnitude of the pulse is closely coupled with the turbine load in the test due to the pulse generation mechanism. On the other hand. and then expands outward quickly as the load increases. Dramatic change of the loops can be seen in the figure: the loop is quite slim at low load. it is convenient to decouple the influence of the turbine load and the pulse magnitude in the 1-D unsteady mode. (a) three different swallowing capacity curves (b) three different mean loads Figure 10. Comparison of hysteresis loops for different swallowing capacity and mean loads 343 . It is clearly demonstrated that the turbine with higher load has stronger ‘unsteadiness’ in the turbine. Figure 10 (b) compares influence of the mean load ( ) on the hysteresis loops under the pulses with the same magnitude of the inlet pressure. However. The for the high load is 16% higher than the original load (middle load) and the 15% lower for the low load. the slope at the mean load point on the curve reduces (subfigure (b)). in terms of the wave action/filling & emptying effect. indicating that the unsteady behaviour. it can be seen that they can be linked to a common factor – the slope of the swallowing capacity curve. as the load increases. 344 . (2) Generally the cycle average efficiency has similar trend as the steady conditions: the nozzled volute is more evident at low pressure ratio thus low load conditions. The smaller slope or higher throttling effect can enhance the unsteadiness of the turbine. It results from the larger absolute flow angle at the nozzled volute exit. but the advantage of the nozzled one is evident at low pressure ratio conditions. thus the slope of the swallowing capacity reflects the throttling effect of the turbine. the peak efficiency shifts to higher velocity ratio with the inducing of the nozzle. As the vane angle reduces. is not obviously influenced by the existing of the nozzle vanes. The maximum efficiency improvement is 11. either by closing the nozzle vane or increasing the turbine load. The unsteadiness can be enhanced by either reducing the swallowing capacity or increasing the turbine load. at 48 kRPM. The slope of the curve. the nozzled volute shows constant efficiency improvement compared with the nozzleless configuration at all operation conditions. it is concluded that the slope of the swallowing capacity curves of a turbine has significant influence on the unsteadiness. the slope of the curve at the same load reduces (subfigure (a)). Moreover. the unsteady behaviour is strongly influenced by the pulse frequency as well as the turbine load. 4 CONCLUSIONS AND DISCUSSIONS Experimental Testing and reduced order unsteady model are applied for the comparison of the performance of the nozzled and nozzleless mix-flow turbocharger turbine under steady and pulsating conditions in this paper. the maximum efficiency improvement by the nozzled configuration is 6. But the magnitude of the efficiency advantage is highly dependent on the pulse frequency and rotor speed. the nozzled turbine shows the advantage of the efficiency over the nozzleless one at most of the conditions. By closely checking those two factors. Instead. at 30 kRPM. which reflects the throttling effect of the turbine. is one of the key factors influencing the turbine unsteadiness.4% achieved at low load 60HZ. (3) The shape of the hysteresis loops of the swallowing capacity for the nozzled and nozzleless volutes are similar. Specifically. Therefore. Actually the effect of the turbine rotor on the flow can be considered as throttling. Conclusions are drawn as following: (1) With the similar swallowing capacity. The unsteadiness can be enhanced by reducing the slope of the curve via either reducing the swallowing capacity (closing the nozzle) or increasing the turbine mean load. which is used to compensate the blockage effect of the nozzle under the constraints of the same swallowing capacity.The results of the reduced order unsteady model further confirm that the unsteady behaviour of the turbine is significantly influenced by the swallowing capacity curve and the turbine load.1% achieved at low load 20HZ. The peak efficiency is similar for two configurations. the loops shift along the steady curves and expand outwards and the covered area increase significantly. or as the turbine load increases. Specifically. (4) The results of the 1-D unsteady model show that the swallowing capacity curve has significant influence on the unsteadiness of the turbine under the pulsating condition. Journal of Turbomachinery.. D. 2000. [6] Romagnoli A. of Mech. and Watterson J K.. A Comparison of Timescales with a Pulsed Flow Turbocharger Turbine.. London. and Martinez-Botas R. A One-Dimensional Study of Unsteady Wave Propagation in Turbocharger Turbines. F. 2009. PhD thesis.REFERENCES [1] Baines N C. and Rajoo S. London. [11] Simpson A T. C. Newton P. Conf. Rosborough R S E. Journal of Turbomachinery. Investigation of Flow in the Nozzleless Spiral Casing of a Radial Inward-Flow Gas Turbine.. and et el. Engrs. [13] Copeland C. 345 . 2008. and Lavy M. [7] Romagnoli A. Conf. 129. GT2000-456. and Rajoo S.. Numerical and Experimental Study of the Performance Effects of Varying Vaneless Space and Vane Solidity in Radial Turbine Stators.. 1969.. Martinez-Botas R F. Martinez-Botas R. F. Martinez-Botas R. [14] Costall A. 10th Int. Radial Turbines-Blade Number and Reaction Effects. Spence S W T. Instn. 2012. 2007.. Martinez-Botas R F. of Mech.. London. Turbochargers and Turbocharging Conference. A Comparison of the Flow Structures and Losses within Vaned and Vaneless Stators for Radial Turbines.. Engrs.135.. Instn. Pro. London. Eng.. and Watterson J K. Engrs. on Turbocharging and Turbochargers. Eng. Turbine Performance Studies for Automotive Turbocharger Part II: Unsteady Analysis. and Martinez-Botas R F. Mixed-flow Turbines for Automotive Turbochargers: Steady and Unsteady Performance. Mech. on Turbocharging and Turbochargers. 9th Int. [2] Bhinder F S. [4] Spence S W T.. Pro. of Mech... 2007.1. Flow in Vaned and Vaneless Stators of Radial Inflow Turbocharger Turbines. 131. 1969. [10] Karamanis N. International Journal of Fluid Machinery and Systems. [3] Rodgers. Spence S W T. Imperial College London. [5] Simpson A T. of Mech. Conf. 2010..132. F. [12] Costall A. Vol. Variable Geometry Mixed Flow Turbine for Turbochargers: An Experimental Study. Mech. on Turbocharging and Turbochargers. Vol. GT2007-28317. Vol. Inst. 2013. ASME paper. 9th Int. 2009.. Turbine Performance Studies for Automotive Turbocharger Part I: Steady Analysis. Martinez-Botas R. ASME Paper. 2007. Instn. Fundamental Characterization of Turbocharger Turbine Unsteady Flow Behaviour. 2009. Vol.. Inst. 2002. [9] Rajoo S. Engrs. A Direct Performance Comparison of Vaned and Vaneless Stators for Radial Turbines. Instn.. Artt D. Vol. [8] Rajoo S. Journal of Turbomachinery.. Unsteady Effect in a Nozzled Turbocharger Turbine. Journal of Turbomachinery. Experimental and numerical investigations on an automotive turbocharger with a transparent bearing section W Köhl Technische Universität Darmstadt, Institute of Structural Dynamics, Germany M Kreschel, D Filsinger IHI Charging Systems International GmbH, Germany ABSTRACT The dynamic behaviour of an automotive turbocharger supported with full floating ring bearings is investigated experimentally and numerically. The rotor orbit on compressor side and, after modifying the turbocharger’s housing, the rotational speed of the floating ring are investigated by the measurement. To determine the ring’s rotational speed, the housing of the journal bearing section has been rebuilt employing a transparent synthetic material. The floating ring on the compressor side is now visually accessible. Using a high-speed camera and a subsequent image analysis in MATLAB, the communication holes around the circumference of the floating ring are detected and the rotational speed can be calculated afterwards. These experimental results are compared with a numerical rotor model. The model simulates a flexible rotor with four translational degrees of freedom and a finite- volume-model of the journal bearings. The numerical results show the subsynchronous limit cycles, which can also be seen in the measurements. The results of the numerically calculated rotational speed of the floating ring and the calculated subsynchronous frequencies in comparison to the experimental results conclude this paper. 1 INTRODUCTION One option to support rotors in automotive turbochargers are floating ring bearings. The floating rings divide the oil film in an inner and outer journal bearing. The advantage of a floating ring bearing compared to a common bearing without floating rings is presented for example in [12], [17] and [18]: the power loss due to friction is reduced and the negative effects of typical phenomena such as oil- whirl and oil-whip are reduced. Numerical simulations, however, show a possible synchronisation of oil-whirl in both oil films which can cause inadmissible high rotor amplitudes. The expression “total instability” is introduced in [16] to describe this synchronisation. These oil induced vibrations are often called constant tone phenomenon. While the rotor vibrates in this state of motion, the subsynchronous frequency changes only barely and the rotor amplitudes increase immediately. The rotor vibrations are transmitted from the bearing system as an undesired sound radiation. In this case, the surrounding parts of the rotor-bearing-system (e.g. turbine and compressor housing, exhaust system) transmit the noise and in worst case depending on the transmission function and their resonance frequencies respectively do behave like _______________________________________ © The author(s) and/or their employer(s), 2014 349 amplifiers. This and other subsynchronous limit cycles occur over a wide speed range of the turbocharger, as experiments presented in [5], [7], [13] and [14] show. The whirl frequencies are a function of the rotational speeds of the rotor and the floating ring. Experimental investigations to determine the rotational speed of the floating ring for validating numerical models of the journal bearings and the rotor model are therefore essential. During the phase of designing the journal bearings, nonlinear models to calculate the fluid forces and friction torques are used to predict critical whirl frequencies, see [1-4], [9] and [12]. Numerical simulations for nonlinear rotordynamics of automotive turbochargers in [4], [6] and [15-18] show the sensitivity of the subsynchronous limit cycles related to the geometry of the floating ring bearing. Basic numerically calculated bifurcation sequences are shown in [16] with varying journal bearing geometry. In general, the rotor is approximated with a finite-element-method (FEM) including gyroscopic effects and a high number of nodes, i.e. a large number of degrees of freedom, see e.g. [13-14]. In some cases the energy equation [3] is solved parallel to the REYNOLDS equation to take the distribution of the oil temperature in the oil film into account. This detailed modelling leads to long computational times (days and weeks). The intension of this paper is to present the experimental results for the rotational speed of the floating ring for validating numerical models. Furthermore it will be shown that a very simple rotor model with very much reduced numerical effort is sufficient to show all relevant effects related to the journal bearings. The numerical results are very well comparable to the experimental results. In section 2, the experimental setup for the modified automotive turbocharger with a transparent bearing section is presented. Section 3 introduces the numerical rotor model of full floating ring bearings. Experimental results and a comparison to the numerical simulations are discussed in section 4. 2 EXPERIMENTAL SETUP To determine the rotational speed of the floating ring, the cast metal housing of the bearing section is replaced by a transparent synthetic material. After designing a 3D-CAD model, a laser lithography process was used to produce the transparent housing. The basic design of the transparent housing is almost identical to the original casted housing. The journal bearing’s geometry and clearances, the connections for the oil supply and the connecting geometry for the turbine and compressor housing are not modified. Deviating from the original casted housing, the surfaces are designed plainly to allow an undisturbed view on the floating ring. The test rig is operated with cold air, which allows the operation of the turbocharger without water cooling. Thus, the water core of the bearing housing was removed. Figure 1 shows the assembled modified transparent bearing section. The test rig is operated with cold gas at a pressure of about 5 bar. A transmission oil (ESSOLUBE X2 20W) is used at room temperature (27 °C). The dynamic oil viscosity is known as a function of the temperature by a previously conducted measurement on a rheometer. Two perpendicular aligned displacement sensors measure the rotor amplitudes of the compressor wheel. A rotational speed sensor is aligned to the blades of the compressor wheel. To avoid aliasing errors, all signals are filtered with an analogue low pass filter. The control system DSpace is used for data acquisition. 350 compressor turbine wheel wheel floating ring Figure 1: Transparent bearing section after machining The area around the floating ring is optically observed by a high-speed camera with monochrome images. The oil flow obstructs the view on the floating ring only slightly. A frame rate of 32,000 fps was chosen to ensure that the communication holes pass the recording area in sufficient time to detect every hole. This setting allows to film an area of 256 x 12 pixels. A sufficiently strong light source has to be used because of the very short exposure time (< 32 μs). An external trigger on the camera allows the synchronisation between the data acquisition using DSpace and the camera shots. All components of the test rig are listed in Table 1. Figure 2 shows the installation of the turbocharger with the transparent housing, the equipment for measuring the rotor orbit and the high-speed camera. Table 1: Components of the cold gas turbocharger test rig Component Type Configuration Turbocharger IHI Charging Systems Waste gate closed International GmbH No compressor housing High-Speed camera Photron: 32,000 fps Fastcam Ultima 512 512 x 12 pixels Light source Shot: cold light 2500 LCD Displacement sensor -Epsilon, Eddy current DT 110-T-S1-M-C3 DSpace Sampling frequency: 8192 Hz Oil ESSOLUBE X2 20W 27°C ≈ 110 mPas oil in air in light source rotational speed sensor floating displacement oil out high-speed ring sensors camera Figure 2: Installation of the turbocharger with the transparent housing in the cold gas test rig, left: side view on the floating ring, right: axial view on the high-speed camera and the sensors 351 In the post-processing, the camera images are analysed in MATLAB. The determination of the floating ring’s rotational speed is basically equivalent with the detection of the changing RGB-colour from black to white, as the drill hole leaves and enters the observed area, respectively. The rotational speed of the floating ring can now be calculated from the number of the drill holes and the time difference between two detected drill holes. For a detailed description of this method see [8]. 3 NUMERICAL MODEL The experimentally determined waterfall diagrams and ring speed ratios presented in section 4 allow to compare the dynamic behaviour of the rotor on compressor side with numerical results. Further on, the ring speed ratio of the floating ring and the rotor rotational speed can now be validated. The complexity of the model is intentionally kept low to show basic effects related to the journal bearing such as oil-whirl. Further, one can compare the numerical calculated critical threshold speeds with the measured ones. Increasing the modelling depth step by step, taking gyroscopic effects into account and using a finite element description of the rotor are scheduled for future investigations. Figure 3 left shows the axial view on the full floating ring bearing. The position of the rotor is given in Cartesian coordiantes and ; the position of the floating ring is given by and . The rotor and the floating ring are rotating with Ω and Ω , respectively. The later used transformation ∗ = + , = + (1) into complex coordiantes (see [2] and [10]) with as the complex number, reduces the number of variables. On the right side of Figure 3 one can see the free body diagram, which is later used to formulate the equation of motion. Ω ℎ Ω ℎ , , , , , , Figure 3: Cut through the full floating ring bearing and free body diagram 3.1 Full floating ring bearing For describing the full floating ring bearing, the following assumptions are introduced: incompressible fluid, constant and equal viscosities in both oil films, constant pressure distribution in radial direction, fully enclosed cylindrical bearings, no drill holes, no defects (e.g. wear), 352 no angular misalignment of the rotor and of the floating ring inside the bearing, GÜMBEL boundary conditions as cavitation model, e.g. [2]. The geometric dimensions of the full floating ring bearing, as it is used for the experimental investigations, require a finite bearing calculation. Therefore, the REYNOLDS equations for the inner oil film 1 ℎ ℎ ℎ + ℎ =6 (Ω + Ω ) +2 (2a) and the outer oil film 1 ℎ ℎ ℎ + ℎ =6 Ω +2 . (2b) are used, to calculate the pressures ( , ) and ( , ) on the rotor and the floating ring, respectively (e.g. [3] and [9]). In the equations (2) and are the coordinates in axial and circumferential direction. The gap function is given by ℎ( , ), is the radius of the bearing and the dynamic viscosity of the fluid. The shear stresses ( , ) are calculated by the assumption of NEWTON’s hypothesis: ℎ ℎ = (Ω − Ω ) + , = Ω − . (3) ℎ 2 ℎ 2 To solve the REYNOLDS equations (2), the finite-volume-method (FVM) is used to discretize both oil films. After discretization of the equations (2) and (3), the resulting forces , , and , , in Cartesian coordinates and friction torques , can be calculated with / / ,, =− , cos , , ,, =− , sin , , / / / (4) , = , , . / 3.2 Rotor-bearing system A simple numerical model of the rotor system is used to calculate the rotor amplitudes and the rotational speed of the floating rings. Figure 4 shows the approximation of the rotor-bearing system. The rotor model contains four mass points on a massless, elastic shaft. The masses and are the masses of the real turbine wheel and the real compressor wheel. The masses , and , in the position of the journal bearings are chosen in a way that the centre of gravity of the rotor is maintained. The two floating rings are approximated by , and , . Each mass point has a translational degree of freedom, which is included in the complex state vectors = , , , , , , = , , , . (5) The two floating rings have additionally one rotational degree of freedom, which leads to the angular matrix =[ , , , ] . (6) 353 com- turbine pressor ⟹ , , , , conical mode S-shape shape Figure 4: Top: approximation of the rotor-bearing system, bottom: two mode shapes: conical mode and first bending mode “S-shape” The natural frequencies of the rotor are known by performed experimental modal analysis on the freely supported rotor. After setting the data for the masses in the matrix , the symmetrical stiffness matrix can be evaluated by matching the calculated natural frequencies from solving the problem det − =0 (7) With 0 0 0 0 , 0 0 = , = (8) 0 0 , 0 0 0 0 with the frequencies known from the experimental modal analysis. The unsupported rotor model without bearings contains two rigid body modes (cylindrical and conical mode shapes) and two bending modes (U-shape and S-shape). A CAD-computer model of the floating rings gives the mass moment of inertia , which is identical for both floating rings in the two journal bearings. Assuming , = , = , the mass matrices for the floating rings can be formulated as: 0 0 floating = , = . (9) 0 0 The complex state vector = , , , (10) the mass matrix = (11) and the stiffness matrix 354 = (12) are now given. Unbalances and the influence of gravity are taken into account using the vectors ∗ ∗ unbalance and gravity , respectively. From integrating the pressures and shear stresses resulting from equations (2) and (3), one gets the acting forces and friction torques of the journal bearing, which are all included in the complex vector ∗ journal bearing after transformation respecting equation (1). The equation of motion can therefore be written as: ∗ ∗ ∗ + = + + . (13) The ordinary differential equation system (ode) of 2nd order (13) is solved numerically after transforming into an ode of 1st order and using MATLAB. The MATLAB solver is designed for numerically stiff problems and a compromise between calculation time and accuracy. 4 COMPARISON BETWEEN NUMERICAL AND EXPERIMENTAL RESULTS The presented experimental results are taken for run downs from Ω , in 2.5 sec to the rotational speed 0 rpm. Therefore, the rotor system is barely affected by disturbances caused by the air flow on the turbine wheel. The radial load is also reduced compared to a run up because the driving air on turbine side is no longer present. The numerical simulations are performed for run downs under the same conditions as mentioned above. Figure 5 shows a comparison of the waterfall diagrams between experiment and simulation for a run down. The diagrams are taken from the measured/simulated amplitude | ̂ | of the compressor wheel. The amplitude | ̂ | is plotted in relation to the maximal radial clearance ℎ , of the full floating ring bearing. The experimentally determined rotational speed of the floating ring is added afterwards. In Figure 5 top, the real rotor system shows two dominant subsynchronous limit cycles. The amplitudes of the synchronous motion can hardly be recognized. The first limit cycle in the speed range (0.1–0.4) Ω , is related to the frequency ≈ 0.48 (Ω + Ω ), which agrees well with the theoretical proposed frequency for the oil-whirl of the inner film, see e.g. [12]. In the speed range (0.7–1) Ω , a second limit cycle occurs. The frequency of this second limit cycle 2 changes barely from 0.41 to 0.44 Ω . This limit cycle produces vibrations with a behaviour of the already mentioned constant tone phenomenon. This oil-whirl state is followed by a jump in the rotational speed of the floating ring. The amplitudes of the inner- whirl are in the range of (1–2) ℎ , ; the resulting amplitudes for the constant tone are up to 1.4 ℎ , . At 0.25 Ω , , the floating ring stops its rotation. The actual bearing geometry of the used turbocharger as presented in Figure 2 is not known without uncertainties. Uncertainties are machining tolerances and material extension due to temperature effects. Therefore, a first simulation with the nominal bearing geometry is performed and compared to the experiment. The synchronous motion due to unbalance is negligible as mentioned above. Therefore, unbalance is not taken into account in the simulation. A dynamic viscosity for the inner and outer oil film = = (27℃) ≈ 110 mPas is used. 355 Calculated waterfall diagram from measurements with the transparent bearing housing maybe bending constant tone 2.5 mode? inner oil-whirl synchronous 2 ȳிோ Ȁȳோǡ௫ ȁݎƸ ȁ௫ Ȁ݄ǡ௫ ͲǤͶͺሺȳ ȳ ሻȀȳ ோ ிோ ோǡ௫ 1.5 1 1 0.75 ௫ ȳோ Ȁȳோǡ௫ 1 0.5 0.5 0 0.25 5 0 0 1 0.75 0.5 0.25 0 ȳிி் Ȁȳோǡ௫ Calculated waterfall diagram from numerical model with nominal bearing geometry outer oil-whirl oil wh whirl 2.5 inner oil-whirl constant tone 2 ͲǤͷȳிோ Ȁȳோǡ௫ ȁݎƸ ȁ௨ Ȁ݄ǡ௫ ͲǤͷሺȳோ ȳிோ ሻȀȳோǡ௫ ȳிோ Ȁȳோǡ௫ 1.5 1 0.75 5 ȳோ Ȁȳோǡ௫ 1 0.5 synchronous 0.5 0.25 5 0 0 1 0.75 0.5 0.25 0 ȳிி் Ȁȳோǡ௫ Calculated waterfall diagram from numerical model using a modified geometry first bending constant tone mode conical mode 2.5 inner oil-whirl 2 ȁݎƸ ȁ௨ Ȁ݄ǡ௫ ͲǤͷሺȳோ ȳிோ ሻȀȳோǡ௫ ȳிோ Ȁȳோǡ௫ 1.5 1 0.75 5 ȳோ Ȁȳோǡ௫ 1 0.5 0.5 synchronous conical mode 0.25 5 0 0 1 0.75 0.5 0.25 0 ȳிி் Ȁȳோǡ௫ Figure 5: Calculated waterfall diagrams, top: experiment with transparent bearing housing, middle: numerical model using the nominal geometry of the bearing, bottom: numerical model using a modified geometry (increased inner clearance , and outer length ) 356 6 ℎ . Comparison between the experimental run down and the numerically simulated run down is carried out using the speed ratio of the floating ring speed and the rotor speed. After determining experimentally the rotational speed of the floating ring using the high-speed camera. While in the speed range of the inner oil-whirl and the constant tone respectively.41 for the numerical model with nominal bearing geometry and the model with a modified geometry. One can recognize a third limit cycle at 0. The rotor enters the constant tone state at 0. .49 and Ω .2. The ring speeds from both numerical models are overestimated. see Figure 4 bottom left. which is closer to the experiment. During this speed range. and an increased outer length is introduced to adjust the subsynchronous motion states closer to the experimental results.49 to 0. . but they are comparable to the experiment. The inner oil-whirl is followed by amplitudes at 0. which has a sufficient correlation to the experimental results. The rotational speed of the floating ring is still calculated too high.65) Ω . To reduce or even to supress the outer whirl in the numerical results. and rotates with amplitudes at 1. The resulting amplitudes of the constant tone are very low. This can be reached by increasing the outer length of the bearing.5–1) ℎ .58 Ω .8 Ω . 357 . . This limit cycle belongs to the oil-whirl of the outer film and does not appear in the experiment.34 … 0. the rotor gets into a conical state of motion. the ring speed ratio results in the relative speed range Ω . However.52 Ω .75 ℎ .55–0. leads to a reduced stiffness of the inner oil film. see Figure 4 bottom right.Figure 5 middle shows the waterfall diagram for a simulated run down using the nominal bearing geometry. One can recognize the oil-whirl of the inner film and the constant tone. The amplitudes of the inner oil-whirl with (0. The frequency of the constant tone alters from 0. the speed ratio will also be decreased. /Ω ≈ 0. The frequency 2 of the constant tone is better estimated with 0. .26 … 0. are lower than the measured ones.56 to 0. The calculated ring speed ratios for the numerical simulations yield Ω . the floating ring speed has to be reduced. the rotor vibrates in the first bending mode shape. /Ω ≈ 0. The outer oil-whirl does not appear anymore within the investigated speed range after modifying the geometry of both journal bearings as mentioned above. with nearly half the frequency of the floating ring’s rotational speed ≈ 0. the calculated rotational speed of the floating ring is higher than the experimentally determined floating ring speed. In general.5 Ω . respectively. see Figure 6. A second numerical model which uses an increased inner radial clearance ℎ . This can be reached by changing the bearing geometry. the stiffness and the damping of the inner and outer oil film have to be modified.1 … 0. It is known from parameter studies that increasing the radial clearance of the inner bearing ℎ . the damping of the outer bearing has to be increased. To adjust the numerical results closer to the experimentally determined rotor movement. Further. the numerical model shows a jump of the compressor’s amplitude before leaving the constant tone in the range (0. The numerical results are presented in Figure 5 bottom.7 Ω . This favours the occurrence of the constant tone phenomenon and the rotor can enter into this state at lower rotor speeds. model 2 provides a better result. By increasing the length . /Ω ≈ 0. . One can find approximate formulas for the ring speed ratio in [3] and [12]. 5 constant tone + outer oil-whirl numerical model nominal bearing geometry 0. This can be improved by taking different viscosities for both oil films into account. Further work will improved the adaptation of the numerically calculated floating ring speed. an important dataset to validate numerical models for rotor systems with floating ring bearings has been created. shows an oil-whirl of the outer oil film.3 constant tone inner oil-whirl numerical model 0.1 0 Ω𝑅 / Ω𝑚𝑎𝑥 Figure 6: Ring speed ratios for the experimental determined rotational speed of the floating ring and the numerically calculated ring speeds Due to a smaller inner clearance. the first numerical model.9 0. although the complexity of the real investigated system exceeds that of the numerical rotor-bearing-model by a multiple. 5 CONCLUSION The use of a transparent bearing housing and a high-speed camera allows to measure the rotational speed of the floating ring on compressor side. the numerically calculated ring speed ratios are overestimated compared to the experiment. However. the viscosity in the inner film gets smaller ( . Investigations to determine the natural frequencies of the unsupported rotor and the oil viscosity with a rheometer improved the quality of the rotor model. 358 .1 experiment 0 1 0. which drives the floating rings. The usage of a constant viscosity in the inner oil film that satisfies < will improve the model of the floating ring bearings. The numerically calculated waterfall diagrams show. Therefore. A simple numerical model of the rotor- bearing-system is presented. Dr. 6 ACKNOWLEDGEMENTS The author thanks Prof. that the threshold speed for entering into the constant tone phenomenon is well estimated. which has not been seen in the experiments.6 0.2 0. ) which is not taken into account for the presented numerical models. is smaller in reality than predicted numerically. Adjustment of the model’s geometric parameters allowed a better match with the measurement results. < .4 0. The friction torque. Experiments for stationary processes and different oil temperatures will provide more data for numerical validation. Furthermore. With this method. that allows to detect the basic effects related to the full floating bearing. that uses the nominal geometry of the bearings.2 modified bearing geometry 0. General agreement with the measurement was achieved.3 0. the amplitudes of the orbit of the compressor wheel during the oil-whirl and the constant tone are comparable to the measured amplitudes. In general. the temperature in the inner film of the real system increases faster.8 0.7 0. 0. Richard Markert for his guidance and contribution to this work and to the cooperation with IHI Charging Systems International GmbH.-Ing.5 0.4 Ω𝐹𝑅 / Ω𝑅 0. STEINHILPER.: Rotordynamics of Automotive Turbochargers. RIVADENEIRA. University of Southampton. 391-397. Auflage.. In: Nonlinear Dynamics 57 (2009). VDI-Fachtagung Schwingungen in Antrieben 2013.1080/10402000701476908 [7] KIRK. B. GJIKA.04. ALSAEED. Springer- Verlag Berlin – Heidelberg – New York. 2006. J. A. A.. P.1016/S0020- 7403(02)00166-2 [2] GASCH. RIVADENEIRA.jsv.. GJIKA. J... R.: Stability analysis of a high-speed automotive turbocharger. VDI- Berichte 2197. In: Tribology Transactions 51 (2008). M. Fachbereich Maschinenbau.1007/s11071-009-9466-3 [16] SCHWEIZER. W.. [4] KAMESH. PFÜTZNER. Auflage.7 REFERENCE LIST [1] ARGHIR. Fachgebiet Strukturdynamik. et al. D... 156-190. doi: 10. In: Journal of Tribology 129 (2007).: Experimental determination of the rotational speed of floating rings in a transparent bearing housing of a turbocharger.. J.. [10] MARKERT. 2119-2132. 509-532. Y. L. In: Journal of Sound and Vibration 321 (2009). In: Journal of Sound and Vibration 330 (2011). In: International Journal of Mechanical Sciences 44 (2002). [11] MUSZYNSKA. R. Springer-Verlag Tokyo. 1035-1046. In: Tribology Transactions 50 (2007). [5] KIRK. CRC Press. R.: Skript zur Vorlesung Rotordynamik. A: Rotordynamics.jsv. NORDMANN. 2013. W.. 427–434..2011. ISBN 978-3- 18-092197-6 [9] LANG.. 2011.: The finite volume solution of the Reynolds equation of lubrication with film discontinuities. R. VDI-Verlag GmbH Düsseldorf. STERLING. KRESCHEL. 343–355. B. 2. doi: 10.: Rotordynamics of small turbochargers supported on floating ring bearings—highlights in bearing analysis and experimental validation. J.. K.: Experimental test results for vibration of a high speed diesel engine turbocharger. oil whip and whirl/whip synchronization occurring in rotor systems with full-floating ring bearings. D. ALSAEED. KORNHAUSER. K. LIPTRAP.1115/1.031 359 . W.028 [17] SCHWEIZER..2436573 [14] SANANDRES. M. O. 422–427.. A. E.: Total instability of turbochargers rotors – Physical explanation of the dynamic failure of rotors with full-floating ring bearings. A. FILSINGER. Faculty of Engineering and the Environment. 2006.2008.: Oil whirl.: Gleitlager – Konstruktionsbücher Band 31. 2. Springer-Verlag Berlin – Heidelberg – New York.: Turbocharger On-Engine Experimental Vibration Testing. GUNTER.2009. Technische Universität Darmstadt. In: Journal of Sound and Vibration 328 (2009).: A virtual tool for prediction of turbocharger nonlinear dynamic response: validation against test data.: Dynamic behaviours of a full floating ring bearing supported turbocharger rotor with engine excitation.. 955-975. 1978. 2005..: Hydrodynamic Lubrication.10. [12] NGUYEN-SCHÄFER. WANG. 2012.1177/1077546309103564 [8] KÖHL. doi: 10. PhD-Thesis. M. In: Journal of Vibration and Control 16 (2010). J.jsv. doi: 10.013 [18] TIAN. doi: 10. ALSAEED. doi: 10. doi: 10. [13] SANANDRES.1016/j. ALSAYED. PENG.1115/1. doi: 10. doi: 10. In: 8. L. H: Rotordynamik.03.1016/j. 4851-4874.: Oil-Whirl instability in an Automotive Turbocharger.: Nonlinear oscillations of automotive turbocharger turbines. et al. J. A. Z.1080/10402000801911853 [6] KIRK. L. NICOLAS. et al. H. R. [3] HORI. B. Springer Heidelberg New York Dordrecht London. Institute of Sound and Vibration Research. 105 – 116. In: Journal of Engineering for Gas Turbines and Power 129 (2006).2464134 [15] SCHWEIZER.. doi: 10.: SIEVERT. R.1016/j. This allows for examining any shape or size of dampers. Miyashita et al. cage and roller) defects on support stiffness and excitation will be examined. (3) and Tanimoto et al. However. 1 INTRODUCTION Turbochargers have commonly been equipped with journal bearings to support the turbines and rotor assembly. The damper will affect not only the turbocharger dynamics but also the bearing dynamics. ceramic balls and usually a machined cage. However. high temperatures and/or corrosive environment. ball bearings have become popular as a replacement for journal bearings in turbochargers. hybrid ceramic bearings are ideal for turbocharger applications. the Reynolds equation is iteratively solved for the squeeze film damper model to determine the damper behavior while accounting for side leakages. The effect of bearing component (inner race. These fundamental characteristics allow for a wide range of performance enhancements in bearing rotor system. lower torque requirement. In this novel approach a six degree of freedom 3D discrete element bearing model was interlaced with a first principle squeeze film damper model to determine the combined stiffness and damping of the turbocharger support. (4) have employed ball bearings in small. Ceramic balls are particularly well suited for use in harsh. Ceramic balls. affecting the bearing life. outer race. 2014 361 . C Lancaster Cummins Turbo Technologies. In addition. stiffer. and electrically resistant. points out that the hybrid ceramic bearing can provide better acceleration response. lower vibrations and lower temperature rise than journal bearings. The combined model was then used to determine the dynamics response of the turbocharger by coupling it with a traditional quasi-static model as well as a time dependent rotor dynamic models. UK ABSTRACT The objective of this investigation is to examine the dynamics of a turbocharger supported by a deep groove or angular contact ball bearing and a squeeze film damper. _______________________________________ © The author(s) and/or their employer(s). as compared to their steel counter parts. automotive turbochargers. USA L Tian. Wang (1). smoother. in his review of ceramic bearing technology. Therefore. corrosion resistant. (2). are lighter. The combined model accounts for the current and the past dynamic states of the system to provide a more accurate support behavior than the current simplified 2-D bearing models used for rotor dynamic analysis.An analytical investigation of turbocharger rotor-bearing dynamics with rolling element bearings and squeeze film dampers A Ashtekar Cummins Turbo Technologies. harder. Keller et al. Hybrid ceramic ball bearings contain steel inner and outer races. (21) and Hendrikx et al. Tiwari (23. the rotor undergoes various mode shapes resulting in complex motion of bearing rotor system. Bou-Said et al. high output turbochargers it is critical to combine the effects of the bearing and shaft/rotor dynamics. BEAST software developed by Stacke et al (26) is known to consider rotor flexibility. fully dynamic discrete element model. (5. investigators have attempted to develop analytical models to study the dynamics of simple rotor systems with rolling element bearings. Prenger’s model included the effect of flexible shafts. In high speed applications. Investigators have attempted to analytically analyze the dynamics of turbocharger rotor system. Meyer et al. Therefore these models are unable to predict the rotor dynamics of turbochargers which use rolling element bearings. Their models consider bearing components as sections of spheres and cylinders. However. (9) demonstrated the advantages of such turbocharger rotor dynamic models by employing them to improve the design of bearings used in a turbocharger. (14) introduced the effects of defects on bearing and demonstrated the vibrations patterns associated with the defects. (17. The model combines a discrete element bearing model and a flexible rotor model to simulate the dynamics of the bearing rotor system. The model developed was capable of analyzing motion of all bearing components. (15) and Ghaisas et al.7) has presented comprehensive models to predict turbocharger dynamics. however they neglected the effect of bearing cage on the dynamics of the system. Inclusion of a complete fluid-film bearing model provided an insight into the effects of bearing dynamics on the dynamics of a turbocharger. San Andrés et al. 24) considered the effects of imbalance and bearing preloading on the rotor dynamics. for a complete understanding and examination of high speed. However. rotor imbalances and operating speeds. however. (16) presented a six degree of freedom. As the bearing size increases. the previous investigators concentrated on the bearing dynamics and ignored the complicated interaction of the roller bearing with the shaft/rotor system. 18) developed a bearing model which included the effects of inclusions. only simple shaft models were considered and the model was unable to handle high speed applications. Bonello (10) implemented non-linear model to study the dynamics of turbocharger on full floating and semi-floating ring bearings. neither the model nor the results are available in public domain. The model was then used to investigate the motion of each bearing components and determine the forces and deflection of the rotor as a function of various operating conditions. 362 . Gupta (11-13) was among the first to present a three dimensional bearing dynamic model. the dynamics of the bearing rotor system becomes critical for comprehensive design and satisfactory operation of the turbocharger. which significantly reduced the computational effort associated with bearing dynamic modeling. In general. In this investigation a model was developed to represent the turbocharger bearing rotor system. most of the work in turbocharger rotor dynamic models has been concentrated on turbochargers with journal bearings.challenges still remain for high speed. (8) also investigated the rotor dynamics of a turbocharger with linear and non-linear aerodynamic bearing models. The results from the model were used to investigate the bearing performance at various preloads. (22) developed a bearing model including the effects of rotor flexibility. however. Pettinato et al. Prenger (25) presented a bearing model capable of modeling tapered roller bearings and angular contact bearings. Saheta et al. however.6. a simplified ideal bearing model was considered and rotor was assumed to be rigid. Sopanen et al. Lim et al. However in their analysis. high output turbochargers which demand large bore bearings operating at DN numbers over 2 million. (19. 20) developed a six degree of freedom bearing model which included the effects of bearing surface defects. Ashtekar et al. Nevertheless. cage dynamics and centrifugal loads were ignored. = + 363 . K is the Hertzian stiffness coefficient. To simplify the overlap calculations. It is possible to consider other shapes in the simulations. between the elements is given by. Rolling element contact forces are considered when balls are in contact with other bearing components. the contact detection schemes become more computationally intensive (Ting (27). the analytical investigation includes a bearing dynamic model which interacts with a flexible turbocharger rotor model to predict the dynamics of rotor system. = + . the turbocharger rotor affects the dynamics of all bearing components. . (29)). =( + )−| − | Where. 2. contact forces and rotor interaction forces are considered as a part of the analysis. rolling elements (balls) and a cage which separates the balls. These bearing components interact with each other directly or indirectly. Although bearing component surfaces deform to some degree when in contact. in the contact model the detailed deformations of the contacting surfaces are ignored and instead the two contacting surfaces are allowed to overlap slightly.1 Dynamic bearing Model A key aspect of modeling the bearing dynamics with Discrete Element Method is obtaining the forces and moments acting on the bearing components. non-conformal contacting solids is given by Hamrock (31) ⁄ 2 = ⁄ ℑ 9 Where R is the curvature sum given by. The turbocharger rotor is supported by the inner race and thus its motion and forces are also affected by the dynamics of the turbocharger rotor. K for two general. This approach of calculating normal contact force is much simpler and less computationally intensive than the method described by Gupta (30). In the current model. the gravitational forces. bearing components are assumed to be made of simplified geometry consisting of sections of sphere and cylinders. Ting et al. The overlap. these deformations are typically very small in comparison to the ball’s characteristic length. Hence. outer race. The degree of overlap is then used to determine the contact forces acting on the bearing components. Matuttis et al. affecting the motion and forces occurring between them. The normal contact force can be determined using the overlap and Hertzian force-deflection relationship ⁄ = Where. however. In this study. As these motions and forces are eventually transmitted from the inner race to all other bearing components. (28). any dynamic instability within the bearing is transmitted to the turbocharger rotor. = + .2 MODEL DESCRIPTION A ball bearing consists of an inner race. The Hertzian stiffness. Similarly. and are the radius of the bodies and and are the position vectors of the respective bodies. The resulting contact forces and moments act equally but in the opposite directions on both of the bodies in contact. = . 2 = 1− 1− + Where . .ℑ= + ln .The values of and are curvatures of the body in X and Y planes respectively. Newton’s second law is used to calculate the linear and angular acceleration of the bodies. which is the difference in instantaneous velocities of bodies in contact.2 Squeeze Film Damper Hamrock’s (31) solution of Reynolds equation is used to calculate the reaction forces due to squeeze film damper on the outer race. Parameters . current study uses the approximate solution provided by Hamrock (31) ⁄ = . Each body has 6 DOF and thus each body is associated with 6 equations that are integrated. =1+ . The moment about the ball center results in rolling motion while the moment about contact ellipse center causes the ball to spin. This tangential force is determined using a traction model. values of A. 2. typically maximum time is reached. . In this investigation. 34). however. Here. . = −1 In addition to a normal force. the Kragelskii’s (32) model is used ( | |) = ( + | |) + . is related to slip velocity. and D were calculated using the method used by Gupta (33). After calculating the total force and moments acting on the bodies. a tangential force exists at the point of contact between the ball and race. 34). System presented in this paper has two bearing and each bearing has 15 such bodies. B. as pointed out by Gupta (30. The above procedure is repeated at each time step using the new component states until some end condition. The tangential friction force is then given by =− | | Evaluation of the tangential friction forces at the contact can be quite involved because of the variations in local slip velocities from point to point in the contact ellipse. The resulting tangential force also creates a moment about the ball center and a moment about the center of contact ellipse. and ℑ require iterative calculations. and the normal force at the contact. for most bearing applications the contact ellipse is sufficiently narrow along the direction of rolling so that the variations in the slip velocity and hence friction force along the semi minor axis can be neglected. are the modulus of elasticity and Poisson’s ratio for the two bodies. However. C. The accelerations are integrated with respect to time to obtain velocities and displacements in linear and angular directions. Thus the total friction force can be evaluated by integrating the friction forces along the major axis of contact ellipse (30. the relative tangential velocity at the point of contact. . For a long damper 364 . is the effective elastic modulus obtained from the elastic properties of bodies in contact and is given by. + = −12 4 This equation does not have an exact solution and thus needs to be solved iteratively. consider the Reynolds equation with side leakages. equation reduces to ℎ + ℎ = −6 = / Thus. the side leakages can be neglected and the Reynold’s equation reduces to ℎ = −12 This can be solved to get the relationship 12 ( + ) =− (1 − ) / The damping coefficient can be expressed as. for shorter bearings. (l/d > 4). is the outer race velocity along Z axis. =r and ℎ = (1 − ) Therefore.assumption. Iterative solutions were obtained for a range of l and c values to generate a database of SFD. The above relationship is suitable for any bearing with l/d ratio greater than four. a solution is evaluated for a bearing position defined by h (or z. is the radius of the bearing outer race. For these bearings. Table 1: Range of l and c values L (mm) 2 3 6 10 100 C (µm) 20 30 50 100 200 365 . ℎ ℎ + ℎ = 12 For a bearing. c is the clearance in the squeeze film damper and z is the position of the outer race CG. is the absolute viscosity of the oil. 12 ( + ) = (1 − ) / Where. Table 1 shows the range of l and c values considered for the study. y in this case). For a given and c. the side leakages can be ignored so that the infinitely long bearing assumption holds true. However. Figure 2 shows the implementation of database into DBM. Figure 1: Database for SFD Figure 2: Short SFD model in DBM 366 . Lower IR motion and reaction forces are primary benefits of well-tuned SFD. the reaction force along Y axis is also calculated.Figure 1 shows the database plots. Higher damping reactions were observed at very small lengths and damping reactions reduced as length increased. after an optimum point the damping reactions shot up as length was increased to approach long SFD assumption. In addition. to include the effects of the anti-rotation pin. However. Both these reaction forces are added to the total forces acting on the OR discrete element. From Figure 3 to Figure 5 it can be seen that the damping coefficient is sensitive to length of the bearing and the clearance. This is primarily due to the geometrical constraints due to larger contact surface between bearing and housing. all outer race rotational degrees of freedom were constrained to be fixed. Similarly. Intermediate values were obtained using linear interpolation. For each case. Please note that the reaction force is opposite to the direction of OR velocity. and Reaction Force at SFD and Damping Coefficient was recorded. Increased clearance had negative effect on reaction forces as well as IR motion. IR motion continued to reduce as the length of bearing increased. Outer Race. DBM was run and the motion of Inner Race. Figure 3: Effect of length on reaction forces Figure 4: Effect of length on damping Figure 5: Effect of clearance on reaction forces 367 . mm per plane. the FRM model is connected to the DBM. the DBM and damper model were used to determine the bearing response which was passed on to the flexible rotor dynamic model (FRM). which governs the dynamics of two ball bearings. the implicit MATLAB® routine ode15s© is employed to solve those ODEs with varying time steps.5 g. The diameter of each beam section is determined from the mass of the corresponding section of the turbine wheel. Due to the highly stiff nature of the ordinary differential equations (ODEs) representing the equations of motion of the rotor-bearing system. though the unbalance at the compressor end is out-of-phase to the unbalance at the turbine end. which consists of 26 nodes as shown in Figure 7.2. Figure 6: Dynamic Bearing Rotor Model In this investigation. it is sectioned to five parts. Figure 6 depicts these interface point interactions as two headed arrows indicating that the exchange of dynamic 368 . The four black arrows show the locations of unbalance. As for the compressor disk. From Node 1 to Node 18 is the turbocharger rotor modeled by Timoshenko beam elements. and the differences of polar moment of inertia and transverse moment of inertia between them are compensated by the additionally lumped inertia properties to the center of each beam section.3 Dynamic Bearing Rotor Model (DBRM) Figure 6 depicts a schematic representation of the bearing rotor system as represented by the DBRM. Through those two interface nodes. Figure 7: 26 Node Dynamic Bearing Rotor Model The two pairs of red triangles mark the locations of center of gravity of the ball bearing inner race. as indicated by the directions of those arrows. The magnitude of the unbalance is 1. The turbine disk is modeled as beam elements from Node 18 to Node 26 with four steps. Thus the dynamic response is passed from one model to other. and their mass and inertia properties are all lumped to the corresponding nodes from Node 2 to Node 6. 62E-09 -1. The results from this method are nonetheless useful to analyze basic steady state dynamics. This matrix can be used as support stiffness for any rotordynamic model of choice to investigate the turbocharger rotordynamics in presence of the REB.05E-05 1. to include the effects of the anti-rotation pin.67E+06 -1.03E-10 1.66E+06 -5.24E+04 -1. run parallel. this method oversimplifies the REB and ignores the internal dynamics and instabilities of the REBs. The two bearings have a single piece outer race. Figure 6 illustrates these linkages shown as lines. Table 2: Bearing Stiffness matrix Bearing Direction Bearing Stiffness Fx 1.05E-05 1.1 Model Interaction Study To allow for the union of an explicit Bearing model with an implicit rotor model. three different methods were used.24E+04 -1.71E-06 1.13E+02 Fx 1.response occurs from both the sides. all outer race rotational degrees of freedom were constrained to be fixed.37E-06 1. The model was subjected to a varying IR motion and the reaction forces from the model were compiled to determine the stiffness matric of the bearing.13E+02 -1.69E-10 Mzx -1.70E-06 -1. Comparison with other methods shows that the basic dynamics can be evaluated to acceptable accuracy. DBM and FRM.23E+04 Fy -5.23E+04 9.88E-08 Myz -9. Finally the single piece outer race of the DBMs is attached to the ground through a spring-damper arrangement representing the squeeze film damper.24E+04 Fy 2.82E-08 -3. Also. the outer races of the two DBMs. communicating with each other at each time step.79E-08 -2.82E-08 1.70E-10 1. namely. 369 .86E-06 -9. However.59E-08 -1.47E-06 1. are rigidly linked to each other.82E-08 -1.13E+02 This is a simple and efficient approach to be incorporated in any rotordynamic model.79E-08 1. In the first method the REB stiffness was evaluated using the DBM.43E+06 -7.23E+04 1.43E+06 1. DBM and the rotor model. 3 ADDITIONAL RESULTS AND OBSERVATIONS 3.86E-06 2.52E-09 -2. The table shows the matrix for a turbocharger bearing. Any motion and/or forces due to turbocharger rotor flexibility affects the dynamics of all the bearing components and similarly dynamic response of bearing components affect the dynamics of the entire turbocharger bearing rotor system.60E-07 1. each representing one of the bearing.38E-06 3.48E-08 1.66E+06 2.24E+04 -7.13E+02 1.88E-08 1.67E+06 2.03E-10 Mzx -1. Therefore.58E-08 Myz 1.48E-08 -3.46E-06 1.82E-08 Turbine Fz 3.23E+04 9.64E-07 2. The two models.60E-09 Compressor Fz 2.43E-09 1. In the third method the DBM was run in parallel to the rotordynamic model. Figure 8: Comparison of REB Models 3. The imbalance affects the rotordynamics as well as has a significant effect on the bearing dynamics. The effects of these bearing instability is examined on the turbocharger.2 Preloading Angular contact bearing are commonly preloaded. however. The forces and displacements are passed back to the rotor model and the simulation continues. Both of these cases are operating at the speed of 50000 rpm with 10 gm-mm imbalance. However. The loss of load between the ball and inner race can cause ball sliding and skidding. This method produces stable solutions and accounts for bearing internal dynamic response. one which has preloading and one without preloading. Each time the rotor model passes on the states of the node. Please note the increased force fluctuation for the case of unloaded bearing. The explicit DBM model is ramped up to the state and allowed to reach a steady state. It is also to be noted that excessive preloading can lead to premature fatigue failure of the bearing. This method does not completely account for the past dynamics of the system but offers an REB solution that is analogues to journal bearing models. Rotor transients have a significant effect on bearings dynamics. the past REB state is used as the starting point and the DBM explicit model is ramped on from the old node state to the new one. the simulation resources and time required are significant. 370 . Hagiu (35) has demonstrated that wrong preloading will cause considerable reduction in bearing life. Figure 9 shows ball loads for two DBRM conditions. However. The results demonstrate the significant effect of wrong bearing preloading in turbocharger. The rotordynamic model is implicit and passes on the node state to the bearing model. these models do have the possibility of diverging solutions in rotordynamic models. This allows for including the transient effects in the model. the model was used with a quasi-static approach model. The results also demonstrate that occasionally the ball-race load becomes zero indicting loss of ball-race contact.In the second method. The results for each of these methods were compared against each other and evaluated for a range of imbalances. e. a defect was introduced in the DBM using the defect models by Ashtekar et al. The analytical model was used to investigate the different approaches to model the REB into the system. Ø = contact angle of the angular contact bearing. a defect on any of the bearing components results in periodic excitation in the system. there is no guarantee that the ball will pass over the defect each time. The model combines a discrete element bearing model and a flexible rotor model. A coupled dynamic model was developed for the ball bearing rotor systems.20) and their effects on the rotordynamics were observed. however. The rolling elements roll over the IR as well as the OR and drive the cage at the same time. they ignore the 3D nature of the bearings. For the analytical investigation rolling element bearing demands significant speed and accuracy of contact force calculations. Figure 9: Unloaded vs Preloaded Bearing Forces 3. An analytical model that meets these demands has been developed. Due to continuous but periodic nature of this interaction. (19. A squeeze film iterative model was also developed to model the squeeze film dampers required to counter the high stiffness of the REB.3 Bearing Defects REB component configuration is of the planetary type with IR being the sun and the rolling elements being the planets. The model differences were highlighted under imbalance conditions to demonstrate the dynamics ignored by 371 . Bd = Ball Diameter. Ball Defect Frequency = (Pd/(2*Bd)) * (N/60) * (1 – (Bd/Pd*cosØ)^2) Cage defect frequency = N/120 * (1 – Bd/Pd*cosØ) OR defect frequency = Nb/2 * (N/60) * (1 – (Bd/Pd*cosØ)) IR defect frequency = Nb/2 * (N/60) * (1 + (Bd/Pd*cosØ)) Where. Nb = Number of Balls. The ball track may be wide and might miss the defect in a periodic manner. 4 SUMMARY AND CONCLUSIONS Investigation into replacing journal bearings of a high speed turbocharger with hybrid ceramic ball bearings requires a detailed analytical model of bearing rotor dynamics. These equations provide a good guidance. i. Pd = Pitch Diameter. This allows for a realistic simulation of the defects and their effects on turbocharger. Thus. N = Speed. 012503. 79-93. 293- 304. Part 2: implementation and results... Kajihara K. Rivadeneria J. 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(25) Prenger N. and Meachum L. “A robust algorithm for ellipse-based discrete element modelling of granular materials. 25-35. p.. 260–265.. 1993. 041103. London. 261-269. and Jacobson B. and Stacke L.. 130(4). 63–71. Purdue University. 224(1). (26) Stacke L.” Wear. (32) Kragelskii I.. 3.” M.” Powder Technology... 1994. “Vibration transmission through rolling element bearings. Advanced Dynamics of Rolling Elements. H Schwarze Institute of Tribology and Energy Conversion Machinery. NOTATION A system matrix q state vector B input matrix T temperature B axial bearing length t time C damping matrix U hydrodynamic circumferential D bearing diameter speed d fluid film damping coefficient u shaft translation degrees of F. f force (vector) freedom x. 2014 375 .Predictions for run-up procedures of automotive turbochargers with full- floating ring bearings including thermal effects and different bearing setups D Vetter. F2 viscosity factors angular direction of fluid film h film thickness force K stiffness matrix lubricant dynamic viscosity K x . the consideration of the hot gas boundary conditions is identified as decisive in order to predict the bearing parameters and consequently the vibrations precisely. Further. the amplitudes of conical vibrational modes are damped. y. Consequently. In particular. The run-up predictions show that the thrust bearing and load have significant influence on the frequencies as well as on the magnitudes of the vibrations. INTRODUCTION Due to light rotor weights and high rotational speeds automotive turbochargers commonly feature full. T Hagemann. the rotor operates above its linear stability limit. As a novel aspect the influence of the thrust bearing and load on lateral vibration phenomena is investigated in detail for this application. The high rotating frequency effects high unbalance forces while static forces are very low. The design with two separated oil films involves the advantage of additional external damping provided by the second film if subsynchronous self-exited vibrations occur and one of the _______________________________________ © The author(s) and/or their employer(s). z cartesian coordinates F0 .or semi-floating ring bearings (FRB/SFRB). F1. Technical University of Clausthal. Germany ABSTRACT Automotive turbochargers feature a variety of boundary conditions influencing the operating behaviour during run-up procedures.Kz turbulence factors density M inertia matrix Φ matrix of eigenvectors p pressure shaft rotational angular velocity 1. In agreement with these authors other researchers identified significant influences on lateral vibrations effected by thrust bearings. g. However. Turbochargers commonly feature hydrodynamic thrust bearings (TB) to balance the axial load. An alternative model is presented by San Andres and Kerth [6] and validated with test data. Concordantly a significant decrease of the ratio between ring and rotor speed is reported if the rotor speed increases. effective clearance. stable limit cycles are reached which enable continuous operation. Porzig et al. Further. [8] show that for hot gas applications the ring-speed and the clearances of the inner and outer film are significantly influenced by the heat flow from the shaft into the inner oil film. [11]-[13].two lubricant films becomes unstable. e. [9] investigated the influence of axial clearance and load on the gyrating movement of the thrust collar and its effect on bending vibrations. [10] studied the lateral vibration of a shaft optionally supported by a thrust bearing at one end. Aside from this function they can strongly influence lateral vibration behaviour of high speed rotating machinery. their nonlinear computational model predicts an excitation of a critical speed by a defect of the thrust bearing. It further includes a validation of the predicted results against measurement data from a hot gas test rig. F1 0 dy . Mittwollen et al. The vibration phenomena reported for turbochargers with FRBs are comprehensively discussed in [2]. The authors demonstrate that an increase of the axial load or a reduction of the clearance can cause a shift of critical speeds and threshold speed of instability to higher rotating frequencies. Instead of a total instability leading to a destruction of the rotor. THEORY Bearing model The prediction of the journal bearing forces generally requires the solution of extended and generalized Reynolds equation F2 p F2 p F1 U x K x x z K z z h x F0 t h . 0 (2) 376 . The influence of the effective bearing geometry on the performance of high-speed turbochargers featuring FRBs is investigated by several researchers [3]-[8]. Trippett and Li [4] developed a thermo-elastic model for the prediction of ring-speed. power loss and local temperature of FRBs and validated it against measurement data. 2. Their results reveal that the thrust bearing provides additional stiffness and damping on the flexible shaft. F2 y F 0 dy . Berger et al. However. This paper presents the influence of the predicted effective bearing parameters and the thrust bearing analysis on the theoretical vibration behaviour of a small automotive turbocharger with FRBs. thrust bearings are neglected in most rotordynamic analyses and no investigations of small turbochargers including thrust bearings can be found. (1) Equation (1) considers the 3D viscosity distribution due to variable temperature in all three space directions of the film by the following factors [14]: h h h dy y y F1 F0 0 . for hot gas applications all parameters strongly depend on the heat disposed from a hot shaft [7]. While synchronous vibrations can be analysed by linear theory [1] the prediction of subsynchronous vibrations requires a nonlinear bearing force model. This model is further applied and verified by measurements of Meyer [5]. Thus. z ) r ( . (4) The forces resulting from both pressure distributions can be superposed in order to predict the fluid film force. If periodicity is fulfilled only one pad has to be analysed and all other information can be generated by rotational transformation. (3). and dij can be listed in non-dimensional form in a look-up table as a function of relative eccentricity ε and attitude angle β.Additionally. Neglecting compressibility of the fluid these to equations are derived assuming the following linearization due to small shaft motion H ( . The attitude angle is shifted between 3° and 6° from one to another radial axis depending on the pad angle. (5) y cos r d21 d22 y Herein and are the lateral velocities between shaft and ring or ring and housing respectively. The direct numerical solution of (1) satisfies the highest requirements on generality and accuracy. the effects of local turbulent flow or Taylor vortices are taken into account by the coefficients Kx and Kz according to the model presented by Mittwollen [15]. In a second step the two first order perturbed equations are solved. Finally. e. the predicted fluid film force is determined by Fx sin r d11 d12 x F Fr . (1) becomes linear and can be separated into F2 pr F2 pr F1 U h and (3) x K x x z K z z x F0 F2 pt F2 pt x K x x z K z z t h . i. in [16] is applied. the transient forces from (4) can be approximated using four damping coefficients dij derived from two perturbed equations. it is time consuming and provides the disadvantage of numerical instability during the iterative solution procedure as the source term on the right hand side implicitly depends on the solution. For every angular position the eccentricity is varied by 25 steps. an approximation procedure for the nonlinear fluid film force proposed by Glienicke et al. z ) X sin Y cos . Consequently. The four non-dimensional damping coefficients d of the fluid film are calculated integrating the solution variables according to 377 . Further. Assuming validity of similitude theory Fr. z ) X Y Y . As rotational speeds are very high it is assumed that the cavitation zone of the film is equal to the one determined for a certain rotor position and exclusive rotation neglecting the time depended source term. X stat stat Herein Π is the film pressure. γr. Fr is the absolute value and γr is the direction of the bearing force resulting from rotation. (6) ( . H is the film thickness and X′ and Y′ are the shaft velocities each in non-dimensional form. On the other hand. In order to approximate the nonlinear force characteristic of the film the distance between two eccentricities is reduced with increasing shaft displacement. At every position of this grid the non-dimensional form of equation (3) is solved based on Elrod’s algorithm [17] and the non-dimensional force due to shaft rotation F r and its direction γr are determined. The linear stiffness and damping coefficients for the collar gyration induced by bending vibration are derived from a perturbation of the thermo-hydrodynamic analysis presented in [20]. Optionally the friction forces can be added to this table and used to predict the ring speed in the rotordynamic analysis. the ring speed is predicted by the time depended friction forces during the run-up procedure in these investigations. However. (7) 0 qk 2 Figure 1 illustrates the procedure to predict the bearing force according to (5) considering the speed dependent fluid film viscosities and bearing clearances. Non-dimensional look-up table Extended Reynolds Equation: Pressure distribution due to shaft rotation F r . 2B / D 1 1 (i 1) cos (i 2) sin dz d 4 B / D 0 d ik .β) First order perturbed equations: Pressure variation due to shaft movement dik = f(ε. γr = f(ε. Rotor model The rotor model consists of a flexible shaft which is modelled by Finite Elements using the Timoshenko beam theory.β) Prediction of dimensional parameters Fr Rotor position and and dik considering effective radial lateral velocities clearance and fluid viscosities Predicted bearing force according to Effective radial clearance and fluid equation (5) viscosities predicted by thermal analysis Figure 1: Bearing force analyses The numerical procedure used to determine the content of the look-up tables is more comprehensively described in [18] and [19]. Compressor and turbine wheel as well as the floating rings are assumed as rigid bodies. For both predictions the ring speed is listed as a function of rotor speed as well. The mathematical formulation leads to a system of ordinary differential equations: 378 . The effective fluid viscosity parameters according to (2) as well as the relative clearances are taken from an a priori thermal analysis according to [4] or [8]. The relation between the physical coordinates and the new modal coordinates ∗ is given by q* Φ1q . The rotor is supported by four-lobe bearings and exhibits a distinct nonlinear behaviour. the gravity-. The eigenvectors of the free system are arranged in a modal matrix Φ which is the basic approach for a modal reduction for example described in [21]. Therefore. The discretized model of the overhung rotor is depicted in Figure 2. The second order system is transformed in a first order state space formulation using a new state vector = ( . the threshold speed of instability slightly decreases if the unbalance increases. The four-lobe bearings are placed at node two and four. . an exact agreement between measurement and prediction cannot be expected. u (8) The vector of forces f contains the unbalance-. 3. The added Figure 2: Overhung rotor model mass as well as the unbalance are located on node six. 379 . Aq Bf(t .and the nonlinear bearing forces. ) and the input vector = ( . For low unbalances typical characteristic of a rotor in journal bearings can be observed. q (9) An eigenvalue analysis is conducted using the homogeneous system of equations. ) . some experimental parameters such as oil supply temperature and the mass moments of inertia of the disk are not defined exactly in [16]. Consistently. The influence of the drive is also not taken into account. A comparison between measured and predicted amplitudes is depicted on the right side in Figure 3. the gyroscopic. Figure 3 shows the predicted amplitudes as a function of the rotor speed for different unbalance values on the left side. There are the first resonance at rotor speeds of about n=4000 rpm and the threshold speed of instability at n=11000 rpm. the good correspondence between the predicted and measured phenomena can be described as very satisfactorily. q) . u. The matrices on the left hand side of equation (8) are constant and time-invariant. The amplitudes of the mean resonance AR show very good agreement. Ku f(t . Though. If the unbalance increases a second resonance arises caused by an oil-whirl which excites the first natural frequency of the rotor-bearing-system. (10) The transformed and reduced system can be integrated by a Numerical Differential Formula (NDF). . VERIFICATION OF THE NONLINEAR ROTORDYNAMIC MODEL The predictions of the nonlinear rotordynamic model are verified with measurements presented for an overhung rotor in [16]. In case of the prediction this value is shifted to u=200 mmg. In the prediction the support structure is assumed as ideally rigid and an interaction with the rotor-bearing system is neglected. Consequently.Mu Cu ) . In the experiment the subsynchronous resonance Asub occurs if an unbalance mass of about u=130 mmg is reached. Whereas the floating ring bearings are positioned at Table 1: Characteristic parameters node seven and eight the thrust bearing is added on node Pa ram e te r Value five.6 / 6. Compressor and turbine Total rotor length lR in mm ≈ 100 wheel are arranged on nodes Total rotor mass m R in g ≈ 100 four and eleven.05 full-floating thrust ring collar S1 thrust collar S2 outer oil film turbine wheel thrust bearing inner oil oil film 1 film compressor thrust bearing wheel oil film 2 Figure 4: Schematic design and rotor model of the turbocharger 380 . RESULTS Figure 4 shows a typical design of a bearing setup for small automotive turbochargers including two FRB and a double acting thrust bearing with two collars.8 listed in Table 1. Further.5 / 11. It is similar to the one investigated in this paper.6 characteristic parameters of the rotor and the bearings are TB diameter D i / D a in mm 8. Figure 4 includes the discretized rotor model. Unbalance u in mmg 0. It is assumed Lubricant 5W40 that unbalances exist on the Oil supply temperature T sup in °C 90 turbine and compressor wheel FRB length B i / B a in mm 3. 200 180 u = 100 mmg 200 300 400 Measurement A R [16] 180 160 Prediction A R Measurement A sub [16] 160 140 Prediction A sub 140 120 Predicted amplitude [µm] 120 Amplitude [µm] 100 100 80 80 60 60 40 40 20 20 0 0 1000 2000 3000 4000 5000 6000 7000 8000 9000 10000 11000 0 50 100 150 200 250 300 350 400 Rotor Speed [rpm] Unbalance [mmg] Figure 3: Measured and predicted amplitudes of the overhung rotor 4.0 / 9. The FRB diameter D i / D a in mm 6.2 with a phase shift of 180°. The rotor consists of a flexible shaft with added masses. The different speed dependent characteristics of the predicted effective film parameters are presented in Figure 6. The results for both thermal analyses are presented consecutively. Then Sub 1 bifurcates into Sub 2 and Sub 3.8 Trippett and Li OF Trippett and Li IF 0. In the speed range from 500 rps up to 1000 rps the unbalance synchronous vibrations and the self-excited vibration Sub 1 exist.4 Porzig et al. Sub 3 is interrupted at a rotor speed range from about n=1500 rps to n=1800 rps. Beside Sub 3 and the unbalance synchronous vibration (Sync) further frequencies can be observed. OF TS Porzig et al.T Sub 3 Sync 2500 Rotor Speed [rps] nFR. IF TS 0. 1 1. OF CS eff 0 Porzig et al.9 Porzig et al. IF TS 0. The measured frequency spectrum of the compressor wheel is depicted in Figure 5.4 500 1000 1500 2000 2500 3000 3500 500 1000 1500 2000 2500 3000 3500 Rotor Speed [rps] Rotor Speed [rps] Figure 6: Predicted viscosities and clearances for the inner film (IF) and outer film (OF) on compressor side (CS) and turbine side (TS) For the modal reduction the eigenvectors up to the fourth bending mode of the free system are used. In case (A) no thrust bearing is arranged. As these frequencies are higher than the ring rotating frequency they are related to whirl phenomena in the inner lubricant film.7 1. 381 .6 Porzig et al. OF TS eff 0 Relative effective clearance / Relative effective viscosity / 0. In case (B) and (C) a double acting thrust bearing is applied and loads of Fth = 22 N (case B) and 87 N (case C) are investigated. This effect can be related to a change of the excited rotor mode by the whirl in the inner oil film.8 0. IF CS 1. OF CS 0. 3500 Sub 1 Sub 2 3000 nFR. IF CS 1 Porzig et al.5 Trippett and Li OF Trippett and Li IF 0. It shows a distinct low frequency vibration at high rotor speeds that can be related to a self-excited vibration in the outer lubricant gap (Sub 3) as it is lower than the measured ring rotating frequency. In the range from 1000 rps and 1500 rps two resonances appear that can be related to Sub 2.C 2000 1500 1000 500 0 200 400 600 800 1000 Frequency [Hz] Figure 5: Measured frequency spectrum at the compressor nose of the turbocharger at run-up according to [8] In order to study the influence of the predicted effective bearing parameters by the thermal a priori analysis and the thrust bearing gyrating stiffness and damping run- up simulations are performed for three different cases.2 Porzig et al.6 Porzig et al. 5 (C) 1500 0 0 0 1000 -0.5 -1 0 500 1000 1500 2000 2500 3000 3500 0.5 500 500 Rotor Speed [rps] 1000 -1 0 0 500 1000 1500 2000 2500 3000 3500 1500 Frequency [Hz] Rotor Speed [rps] Figure 7: Predicted frequency cascades and time signals at the compressor nose using the thermal analysis according to Trippett and Li [4] 382 . Here. Sub 3 is only slightly present if no thrust bearing is considered (case A). Beside the unbalance synchronous vibration the predicted frequency spectrum in Figure 7 shows two subsynchronous vibrations that can be related to the inner oil film and are called Sub 1 and Sub 2.5 1 3000 -1 Amplitude 2500 0 500 1000 1500 2000 2500 3000 3500 0. If thrust force increases the following effects on Sub 1 can be detected: the beginning and ending are shifted to higher rotor speeds the amplitudes decrease The influence of the thrust bearing on Sub 2 and the unbalance synchronous vibrations are insignificant. Sub 3 case (A) Sub 2 case (B) Sync 3500 3500 1 3000 1 3000 Amplitude Amplitude 2500 2500 0. With further rising rotor speeds the frequency of Sub 2 is not changing anymore but rather stays constant.5 0. Sub 1 starts at rotor speeds of about n=100 rps. This effect is well-known as locking effect if the oil-whirl turns into the oil-whip at a natural frequency of the rotor-bearing system.5 2000 0. Its frequency rises from f=60 Hz at the beginning to f=220 Hz at a rotor speed of n=650 rps.5 2000 2000 1500 1500 0 0 0 1000 0 1000 500 500 500 500 Rotor Speed [rps] Rotor Speed [rps] 1000 1000 Sub 1 1500 0 1500 0 Frequency [Hz] Frequency [Hz] case (C) 0.Results using a priori thermal analysis according to Trippett and Li [4] The turbocharger run-up starts at a rotor speed of n=20 rps and is increased linearly up to 3500 rps in 6 s.5 0 (A) -0. The subharmonic vibrations caused by a whirl in the outer lubricant gap will subsequently be called Sub 3. At this rotor speed Sub 1 bifurcates into Sub 2 which starts with a vibration frequency of 320 Hz. Sub 2 reaches a frequency of f=690 Hz at the end of the run-up procedure.5 Amplitude (B) 0 3500 -0. case (A) case (B) Sub 3 Sub 2 Sync 3500 3500 1 3000 1 3000 Amplitude Amplitude 2500 2500 0. For maximum thrust load case (C) the beginning of Sub 3 shifts back to lower rotor speeds of n=1000 rps.5 1 3000 -1 Amplitude 2500 0 500 1000 1500 2000 2500 3000 3500 0. In case (A) Sub 3 starts at a rotor speed of n=650 rps with a vibration frequency of 80 Hz and reaches a frequency of 310 Hz at the end of the run-up simulation. Compared to the above presented predictions Sub 3 exists over a wide range of rotor speeds and dominates the vibrational amplitudes as depicted in Figure 8. At light thrust loads of case (B) the beginning of Sub 3 is delayed to higher rotor speeds of about n=1500 rps and the amplitudes are reduced.Results using a priori thermal analysis according to Porzig et al. [8] 383 .5 0 (A) -0.5 2000 0.5 0 1500 0 (C) 0 1000 -0.5 Amplitude 3500 0 (B) -0.5 0. Additionally.5 2000 2000 1500 1500 0 0 0 1000 0 1000 500 500 500 500 Rotor Speed [rps] Rotor Speed [rps] 1000 1000 Sub 1 1500 0 1500 0 Frequency [Hz] Frequency [Hz] case (C) 0.5 -1 0 500 1000 1500 2000 2500 3000 3500 0. particularly at lower rotor speed the start frequency of Sub 1 remains nearly constant the end of Sub 1 is shifted to lower rotor speeds for light thrust loads and back to higher rotor speeds for high thrust loads the frequency at the end of Sub 1 first increases and then decreases the maximum amplitude of Sub 1 slightly decreases Again the influence of the thrust bearing and load on Sub 2 and the unbalance synchronous amplitudes are insignificant. [8] The same run-up simulations are conducted using the effective bearing parameters determined by the theoretical results of Porzig et al.5 500 500 Rotor Speed [rps] 1000 -1 0 0 500 1000 1500 2000 2500 3000 3500 1500 Frequency [Hz] Rotor Speed [rps] Figure 8: Predicted frequency cascades and time signals at the compressor nose using the thermal analysis according to Porzig et al. the following statements can be derived if a thrust bearing is arranged and thrust load increases: the amplitudes of Sub 3 decrease. [8]. [8] (right) 5. As the heat flow from the shaft into the inner oil film is considered the simulated temperature distribution differs from the one predicted by the first analysis. This effect can be related to the very large amplitudes of the subharmonic vibration at rotor speeds higher than n=2000 rps. in additional investigations similar tendencies were observed for phase shifts between the two unbalances. the dominant amplitudes of the self-excited vibration induced by the outer film are only predicted based on this data and are experimentally identified in [8]. CONCLUSIONS The presented results show that the effective parameters of the lubricant gap and the arrangement of a thrust bearing have significant influence on the predicted results of run-up simulation. The increased eccentricities in the outer lubrication gap induce high frictional forces which decelerate the floating ring. The ring speed is underpredicted compared to the measurement data. The ring speed predicted using the thermal model according to Porzig et al. The correspondence between prediction and measurement data was remarkably improved using the a priori thermal analysis which considers the hot gas boundary conditions. The correspondence between measurement and prediction is significantly improved. thermal model according to Trippett and Li [4] (left) and Porzig et al. Nevertheless. The consideration of thrust bearings especially influences the prediction for self-excited vibrations that are typically related to conical modes. In particular. the relation between the clearances of the inner and outer lubrication gap is changed and the ring speed is modified. 384 . The results of the prediction in case of the thermal model according to Trippett and Li [4] are depicted in Figure 9 (left). 700 700 600 600 Floating Ring Speed [rps] Floating Ring Speed [rps] 500 500 400 400 300 300 200 compressor side prediction 200 compressor side prediction turbine side prediction turbine side prediction compressor side measurement [8] compressor side measurement [8] turbine side measurement [8] turbine side measurement [8] 100 100 500 1000 1500 2000 2500 3000 3500 500 1000 1500 2000 2500 3000 3500 Rotor Speed [rps] Rotor Speed [rps] Figure 9: Measured and predicted ring speeds by comparison. Despite rising rotor speeds a decrease of the compressor-sided floating ring speed is observed in the measurement. Consequently. [8] and a dynamic torque balance is depicted in Figure 9 (right). The results indicate that a precise modelling of the speed depended thrust load and its interaction with the thrust bearing can be a key feature improving run-up simulations for this application. The predictions were performed for an estimated unbalance configuration.Ring speed The measured and predicted ring speeds are presented in Figure 9. B. London. on Turbochargers and Turbocharging. Prof.” ASME Paper No. Bhattacharya. “Effect of Hydrodynamic Thrust Bearings on Lateral Shaft Vibrations”.. [11] Lie.. N. K. pp.. 87- GT-110. B. “A Virtual Tool for Prediction of Turbocharger Nonlinear Dynamic Response: Validation Against Test Data”. (2009). An advantage of the presented analysis is that it is highly time-effective and run-up simulations can be performed in less than two hours on a common desktop pc. pp... gleitgelagerte Rotoren”. 156-190 [3] Born. 113. However.. Part J: J... (1983). D. R. the hydraulic interaction between the two films is neglected and the transient pressure built-up is not taken into account for the prediction of the floating ring speed. C. pp. Li. K. Raetz. of Engineering for Gas Turbines and Power. R. Conf. Germany [6] San Andrés. 102507 [8] Porzig. 134(10). J. J. (1991). 328(1-2). Proc. (2014). Raetz. L.. “Äußere Lagerdämpfung für sehr hochtourige. Chinta. there are some assumptions that have to be discussed critically. Glienicke. G. V. “Influence of Axial Thrust Bearing Defects on the Dynamic Behavior of an Elastic Shaft”. Bonneau. 218. J.. Eng.. Inst. Further. H.. 2004. 1–14. TH Karlsruhe. [4] Trippett.. “High-Speed Floating-Ring Bearing Test and Analysis”. ACKNOWLEDGEMENT The Authors would like to thank Mr H.” Proc. Hegel. [5] Meyer. “Thermal Effects on the Performance of Floating Ring Bearings for Turbochargers. “Analytical and Experimental Investigation of the Stability of the Rotor-Bearing System of a New Small Turbocharger. Schwarze.... 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N. (2013).. B. 4. VDI-Verlag... T. R. “Nichtlineare Rotordynamik”. “Strukturdynamik: Diskrete Systeme und Kontinua”. Dusseldorf. “A Generalized Reynolds Equation for Fluid Film Lubrication”. Mechanical Sciences. T.... Schwarze. 1981. 897-908 [14] Dowson. 10th EDF/Pprime Workshop. “Measurement and prediction of the dynamic operating conditions of a large turbine tilting-pad bearing under high circumferential speeds and heavy loads”. K. Kraft. Eilers. M.. pp. (2012). [13] Chang-Jian. “Betriebsverhalten von Radialgleitlagern bei hohen Umfangsgeschwindigkeiten und hohen thermischen Belastungen – Theoretische Untersuchungen”. GT2013-95074. (1996). pp. B. France [21] Gasch. C. Dettmar. (1993).[12] Lie. pp. [16] Glienicke. R.-W. 112-120. further improvement of fuel consumption is needed in automotive engines. turbo-charging is widely applied with the concept of downsizing which is to reduce engine capacity while preserving performance. In this study. It is validated that CFD results can be used for qualitative evaluation of excitation forces. a variable flow capacity turbine which has variable nozzle vanes is mainly used in automotive diesel engines currently. and compared with experimental data. due to the demand of CO2 reduction. unsteady CFD analyses are conducted to investigate flow mechanism that causes aerodynamic excitation forces. Kawakubo (1) conducted unsteady CFD analysis of the variable nozzle turbine where the pressure ratio is changed with fixed nozzle opening to evaluate the flow mechanism of the impeller _______________________________________ © The author(s) and/or their employer(s). it is needed to predict the aerodynamic excitation forces at the design. Then resonant vibration level is calculated by the CFD and FEM. Several researches for unsteady rotor-stator interaction within radial turbines have been published and these evaluated the vibratory response quantitatively by estimating excitation force using the unsteady CFD analysis. As a result. In order to reduce the fuel consumption. Therefore. Japan ABSTRACT It is important to predict the resonant vibration level for the design of a variable nozzle turbine. it is found that excitation forces are affected by shock wave and clearance flow and can be modeled with a linear function of impeller inlet dynamic pressure and Mach number. H Hattori IHI Corporation.A study on unsteady aerodynamic excitation forces on radial turbine blade due to rotor-stator interaction W Sato. NOMENCLATURE VGS ：Variable Geometry System VPF ：Vane Passing Frequency CFD ：Computational Fluid Dynamics S/S ：Suction Surface P/S ：Pressure Surface L/E ：Leading Edge T/E ：Trailing Edge 1 INTRODUCTION In recent years. To satisfy the requirement of such engine performance for the whole operating range. 2014 389 . On the other hand. the unsteady interaction between the nozzle vanes and the impeller blades generates a periodical unsteady loading which excites the impeller blade vibration. by applying the variable geometry system. A Yamagata. parallel computing using PC cluster was applied for this unsteady calculation. Schwitzke et al. the total number of grids became 16 million for the nozzle and the impeller passage. which were nozzle closed. The number of nozzle vanes and impeller blades are 14 and 9 respectively. Computational domain includes only a nozzle and an impeller passage and an inlet scroll was not included. (7) also conducted unsteady CFD analysis and showed three flow mechanism which is responsible for the blade excitation could be identified.2 Numerical Procedure In this study. However. and a rotating part. An H-type structural grid is used for each blade passage. the actual turbine has clearance gaps for both at the hub and the shroud side. and to clarify each flow mechanism. Figure 1(b) shows an axial view of the nozzle vanes and impeller blades. which was developed for the passenger vehicle's diesel engine in IHI. total pressure and swirl angle at nozzle inlet were estimated from 1D calculation of an inlet scroll. the MUSCL-ROE TVD scheme was used. The first is rotor-stator potential interaction. The impeller diameter is about 50mm. resonant vibration level was calculated by the CFD and FEM. and compared with experimental data. In order to reduce the computing time. Therefore. For turbulence closure one-equation Spalart-Allmaras model was used. As inlet boundary conditions. The impeller rotating speed was set to match the resonant frequency. impeller. This is because the flow distortion due to the inlet scroll's asymmetry is rectified and suppressed within a nozzle passage (3). CFD analyses and vibratory measurement were conducted for three nozzle openings. High order accuracy can be achieved by the MUSCL scheme and Shock wave can be captured by the TVD scheme.excitation force. It is necessary to use the full annulus model for unsteady CFD analysis because there is no cyclic symmetry. medium and fully-opened conditions and also conducted at different pressure ratio conditions for each nozzle opening. and the third is a separation vortex on S/S near impeller L/E. which are inlet scroll and nozzle. these gaps for shroud and hub sides are set equally and its amount is set to 1. The turbine is composed of stationary parts. there are not so many papers that systematically describe the effects of nozzle opening and turbine pressure ratio on the turbine impeller excitation force. The minimum size of wall adjacent grid is set as y+ value to be lower than 3 in order to resolve a viscous sub-layer sufficiently. the second is wakes of vanes. Figure 1(c) shows meridional view and axial view of the computational mesh. He showed the excitation force at the impeller leading edge became higher than other region due to the leakage flow at the nozzle side clearance and the shock wave generated near the nozzle trailing edge. Since the computational region could not be reduced anymore. 2 METHODOLOGY 2. After that. In this study. 2. It was validated that the CFD could capture excitation forces qualitatively. For the convective fluxes. Finally. unsteady 3D Reynolds-averaged compressible Navier-Stokes equations were solved using the in-house CFD code developed in IHI Corporation. the authors conducted unsteady CFD analyses for the variable nozzle turbine to investigate the effects of nozzle opening and pressure ratio on excitation force.1 Turbine Configuration Figure 1(a) shows a cut-way view of the RHF5V turbocharger used in this study. In order to ensure the smooth operation of variable nozzle vanes.5% of the nozzle passage height. and 3 weeks computational time is needed to obtain periodically converged solution for each case. In this analysis. which is calculated from the impeller blade's eigen 390 . principal aerodynamic parameters for the excitation force were identified and the excitation force was modeled with a simple linear function. optical sensors direct the light to the blade tip and detect the reflected light to register the blade pass timing.3 Experimental Procedure Figure 1(d) shows the procedure of non-intrusive vibration measurement system. Calculation cases investigated in this study are summarized in Table1. An unsteady pressure difference on each computational grid point is analysed by Fourier transform and 1st VPF component of excitation force is extracted to use for evaluating strength of the excitation force. 2. the inlet pressure was changed at the constant outlet pressure. Relative motion from the un-deformed position is calculated by the delay or the progress of blade passing time. The impeller interacts with the flow from the nozzle as shown in Figure 2(a) and the impeller is excited by an unsteady pressure difference between P/S and S/S as shown in Figure 2(b). and vibratory response is determined.1 CFD model and measurement system 391 .frequency and the number of nozzle vanes. In this system. Nozzle Vane (VGS) Nozzle Vane (VGS) Open condition Medium condition Closed condition Scroll Impeller Impeller (a) Cut-way view (b) Axial view of CFD model (c) Computational mesh Detecting blade Optical sensor vibratory motion θ2 Data acquisition & Analysis system θ1 Vibratory response 1/rev sensor (d) Measurement system Fig. Figure 2 shows the procedure of calculating the excitation force from the unsteady CFD results. When the pressure ratio was changed. From this result. which are the medium pressure ratio. we can derive that the impeller excitation force at nozzle opened condition is proportional to the impeller inlet total density. Figure 4(b) shows the variation of area-averaged excitation force near the impeller L/E for the different nozzle opening and the pressure ratio to evaluate the effect of different pressure ratios quantitatively. Figure 4(b) below chart shows the excitation force normalized by the dynamic pressure based on the impeller inlet stagnation density and the impeller peripheral speed. In the case of the open condition. Pressure ratio. In this study.2 Effect of different pressure ratios Figure 4(a). At the open condition. Figure 4(b) above chart shows that at every nozzle opening the excitation force becomes higher with increasing pressure ratio and increasing rates depend on the nozzle opening. the flow is choked and the calculation under the high pressure ratio was unstable and this result could not be achieved. efficiency and mass flow rate was normalized by the representative values. (b) shows the distributions of 1st VPF component of excitation force on the impeller blade calculated from the unsteady CFD result. respectively. At each nozzle opening. the excitation force becomes higher with increasing pressure ratio but the excitation force distribution is almost unchanged. On the other hand. the normalized excitation force is almost the same and not affected by the difference of pressure ratio. 3.1 Turbine performance Figure 3 shows turbine performance characteristics predicted by unsteady CFD calculations. Table1 Computational cases Nozzle opening Pressure ratio Case1 closed low Case2 closed medium Case3 closed high Case4 medium low Case5 medium medium Case6 medium high Case7 opened low Case8 opened medium Normalized excitation Normalized loading FFT force Rotation Normalized time Excitation order (a) Interaction (b) Unsteady loading of (c) FFT analysis between nozzle vanes the impeller blade and the impeller Fig. the impeller excitation force under the same pressure ratio becomes the highest at the medium nozzle opening.2 Procedure to derive an unsteady excitation force on the impeller 3 RESULTS AND DISCUSSION 3. These effects of different pressure ratios at different nozzle openings are discussed in the following sections. the normalized excitation forces at the closed and medium conditions become higher with increasing pressure ratio. 392 . the maximum efficiency and the maximum mass flow rate. 0.0 1.0 1.1 Closed Medium Open condition condition condition 0 0.0 -0.6 0.4 0.2 1.2 1.2 Normalized pressure ratio 1.4 Normalized pressure ratio (a) Contour of meridional view (b) Area-averaged excitation force near impeller L/E Fig.6 0.1 Normalized efficiency 0.4 (Nondimensional) 0.8 1 1.8 Closed Medium Open condition condition condition 0.0 0.6 0 0.6 0.2 Low 0.2 -0.2 1.8 1 1.3 Turbine performance characteristics predicted by unsteady CFD Open condition (Dimensional) Medium condition High MAX Closed condition 30 MIN Excitation force (Dimensional) 20 Pressure ratio 10 0 Normalized excitation force 0.2 Normalized mass flow rate Fig.2 0.4 0.2 0.1 -0.8 1.8 1.3 -0.30.6 0.4 1st VPF component of excitation force 393 .4 0 0.4 1. 0 Low pressure ratio Medium pressure ratio Absolute Mach number (a) Open condition 1.2. and zero means the flow is incompressible. therefore normalized excitation force is considered to become the same. At the lower pressure ratio. 3. not only the density (the dynamic pressure) but also the shock wave due to high Mach number has strong effects on the excitation force at the nozzle closed and medium conditions. and positive and negative values show the decrease and the increase of density. At the opened condition. respectively.4 + 0 0. Strong pressure disturbance due to this shock wave at the downstream of nozzle generates the pressure fluctuation when the impeller blades are passing the nozzle vanes and causes the strong impeller excitation force.0 - High Shock wave Pressure ratio Low Closed Medium Closed Medium condition condition condition condition Absolute Mach number Dilatation shock (Divergence of velocity vector) (b) Closed and medium condition Fig. but at the medium pressure ratio. Therefore. the results at the nozzle closed and medium conditions are discussed. Definition of the dilatation is the divergence of the velocity vector. the region where Mach number exceeds unity exists at the downstream of the nozzle and spreads by the increase of pressure ratio. Mach number distributions are unchanged. and a clearer shock wave stands at high pressure ratio. Figure 5(a) shows the instantaneous Mach number distribution at mid-span. 1.At first the result at the nozzle opened condition is disscussed. Figure 5(b) shows the dilatation contours to evaluate the strength of the shock wave occurred near the nozzle exit. weak shock wave stands near the nozzle T/E. no clear shock waves are observed.4 0. From Figure 5(b).5 Instantaneous Mach number and dilatation distribution at mid-span 394 .1 Effect of shock wave Next. the flow is choked near the impeller T/E and therefore Mach number distributions between nozzle vanes and impeller blades shown in Figure 5(a) are unchanged although pressure ratio was varied. 4 Flow impinges Low pressure ratio on impeller 0.2). which can simplify the preliminary design process.1 2 2 Closed Medium 0. though the unsteadiness at the impeller L/E is relatively lower due to the relatively lower nozzle loading.2 Effect of nozzle clearance flow Figure 6 shows normalized excitation force near impeller L/E and effect of nozzle clearance flow.3.1 0. When out of the clearance flows (the circumstance the blade 2 is going through).0 condition condition 0. Figure 6(b) shows velocity vectors (showing absolute velocity vectors in the stationary frame and relative velocity vectors in the rotating frame).5 High pressure ratio Medium pressure ratio 0.3 Low 0.4 0.6 Normalized excitation force near impeller L/E and effect of nozzle clearance flow 3. the absolute flow direction is more radially-inward so the flow impinges on the S/S and causes positive unsteady pressure on the S/S. On the basis of this fact.2. 395 .0 0. Absolute Normalized excitation force 0. the excitation force is affected by not only the density (the dynamic pressure) but also the shock wave and nozzle clearance flow.0 shock shock span from hub Velocity vector edge line on the impeller S/S Medium condition under low pressure ratio condition (a) Normalized excitation force (b) Instantaneous velocity vector distributions near impeller L/E distribution near hub endwall Fig.3 0.5 High pressure ratio High 0. A similar explanation may apply to the medium condition. Effect of the impingement is stronger when pressure ratio and nozzle vane loading is high.8 1. In a nozzle clearance flow (the situation the impeller blade 1 in Figure 6 is experiencing). so the flow does not impinge on the S/S. the flow is well guided by the nozzle vanes.0 span from hub Closed condition Normalized excitation force 0.2 Pressure ratio 0.8 1.0 0. this section aims at deriving a model function to predict excitation forces using aerodynamic parameters of preliminary turbine design.0 0.3 Correlation between Mach number and normalized excitation force As described in the previous section (3.6 0.1 and 3. These two alternating flow conditions create a highly unsteady loading at the hub and shroud section near L/E as shown in Figure 6(a).4 Medium pressure ratio Low pressure ratio Relative 0.2 0.6 0.2 0.4 0.2 1 1 0. Effect of shock wave and nozzle clearance flow become stronger when nozzle vane loading and Mach number at the nozzle exit are high. ANSYS was applied as a FEM solver. On the other hand. and vibration response is low. And the CFD was validated by comparison of predicted and measured response. Mach number and normalized excitation force increase with increasing pressure ratio. the CFD is considered to capture the trend of the vibration response due to nozzle opening change and pressure ratio change.3). 4 CONCLUSIONS In this study. Figure 7 shows correlation between Mach number and normalized excitation force.4 Comparison of predicted and measured response Finally the CFD results are validated by the comparison of predicted and measured response. If nozzle clearance on the shroud side in the measurement is larger than that in the CFD. the unsteady CFD analysis of the variable nozzle turbine was conducted to investigate the effects on the impeller excitation force of nozzle opening and pressure ratio. 396 . and the prediction shows about half value of the measurement under the medium condition (Gradient of the prediction curve is 0. This is because it is considered that time-averaged impeller incidence angles of the closed and medium condition are close to each other and therefore the flow patterns at the closed and medium condition are relatively similar. this methodology will be applied on other turbine models to investigate the validity of the derived correlation. but also with one dimensional steady calculation at the preliminary design of a turbine. 3. Figure 7 shows correlations at closed and medium conditions are almost the same and could be approximated by a linear function.6 and that of the measurement is 1. It is found that Mach number as shown in Figure 7 has a stronger correlation (a stronger linearity) with the normalized excitation force as compared to the pressure ratio as shown in Figure 4(b). As a future work. Figure 8 shows comparison of predicted and measured response. With regard to the sensitivity of pressure ratio on vibration response. This is because it is considered that time-averaged impeller incidence angle of the open condition is negative although that of closed and medium condition is positive and similarly the flow pattern of the open condition is quite different from that of the closed and medium condition. Prediction is conducted by the unsteady CFD and FEM structural analysis. From these results. The conclusions were derived as follows. This is possibly because nozzle clearances in the measurement are a little different from that in the CFD. Figure 6(a) shows that excitation force near the impeller L/E tip of the CFD result is small under the medium condition. the excitation force is proportional to the impeller inlet density with varying pressure ratio because the flow is choked near the impeller T/E and Mach number distributions were the same. At the open condition. Both the vibration responses of prediction and measurement are normalized by their maximum values. the prediction shows good agreement with the measurement under the closed condition. It is showed that the excitation force can be modeled with a simple linear function of one-dimensional aerodynamic parameters (dynamic pressure and Mach number) at the closed and medium condition. Therefore aerodynamic excitation forces can be simply predicted not only with three dimensional unsteady CFD. which lead to higher excitation force. Therefore modal force at the L/E tip is considered to be low although modal displacement at the L/E tip is large. the sensitivity (gradient of the curve) of the prediction increases with decreasing nozzle opening similarly to the measurement. effect of nozzle clearance flow becomes stronger and could lead to high sensitivity of pressure ratio on vibration response. The correlation of the open condition is quite different from that of the closed and medium conditions. At the closed and medium condition. 8 1.6 0.0 1.8 Medium condition 1.4 0.8 1.6 excitation force 0. The excitation force could be modeled with a linear function of one-dimensional aerodynamic parameters (dynamic pressure and Mach number).4 0.4 Normalized pressure ratio Fig.0 1.2 Open condition Closed condition Medium condition 1.7 Correlation between Mach number and normalized 0.4 0.2 0. effect of the shock wave and the clearance flow near the nozzle T/E becomes stronger when the pressure ratio and Mach number is high.3 0.2 Measurement Normalized vibration response 0.8 0.2 1. 1.6 0.4 0.4 1.0 1.2 0.At the nozzle closed and medium condition.8 1.1 1.0 0.6 Norm'd pressure ratio=1.0 Closed condition 0. Due to this knowledge.2 0.8 Fig.8 Comparison of predicted and measured response 397 .6 0.2 Prediction Normalized excitation force 1.8 1. The CFD was validated by the comparison of predicted and measured response.0 0.8 0.2 1.0 0.0 nozzle outlet 0.8 0.2 0.0 0.6 0.8 0.4 Open condition Absolute Mach number at 1.0 1.2 0.2 1.0 0.0 1.0 0. excitation forces can be simply predicted at the preliminary turbine design. The CFD could capture the trend of the vibration response due to nozzle opening change and pressure ratio change.2 0.6 0.2 1.0 1. . ASME-GT2008-50461. S. “Unsteady rotor-stator interaction of a radial-inflow turbine with variable nozzle vanes”.. 2012. (4) Tamaki. AICFM9-224. H. Ono H. 2008. T. ASME-GT2010-23677.. Int. 2010.. “Numerical analysis of aerodynamic excitation of blade vibrations due to nozzle guide vanes in radial inflow turbines”. (6) Kulkarni. 2012. 2008. Yamagata. (7) Schwitzke. (2) Ota.. 2007. (5) Hattori. Schulz. S. “A numerical study on the three-dimensional flows in the scrolls and nozzles of a radial turbine for a variable geometry automotive turbocharger”.. Bauer. A. T. and LaRue. and Masakazu. G. “A study on unsteady aerodynamic excitation forces on radial turbine blade due to rotor-stator interaction”. H. 2007. Masaru U. Gas Turbine Congress 2007 Tokyo.. TS-028. A. Masaru. U. Turbomachinery. A. ASME-GT2008-51355. (3) Ota.40.REFERENCES (1) Kawakubo.. H...78-84. ISROMAC-14. “Mistuned vibration of radial inflow turbine impeller”. pp. 398 .. “The effect of clearance flow of variable area nozzles on radial turbine performance”. G. H. “Vibratory response characterization of a radial turbine wheel for automotive turbocharger application”. Shinya.-J. Vol. M.. and Unno M.. and Kawakubo.. Proc. The performance parameters include aerodynamic objective at multi operating points such as efficiency and mass flow factor as well as stress and vibration parameters and turbine inertia. NOMENCLATURE ANN artificial neural network a polynomial coefficient CFD computational fluid dynamics α1 blade leading edge angle DOE design of experiment α2 blade trailing edge angle DRVT blade loading d(rVθ)/dm W1 inlet passage width EA evolutionary algorithm W2 outlet passage width FEA finite element analysis x design parameter LE leading edge y performance parameter TE trailing edge ^ normalised value MOGA multi-objective genetic algorithm GA genetic algorithm RBF radial basis function RSM response surface method _______________________________________ © The author(s) and/or their employer(s). University College London.A 3D inverse design based multidisciplinary optimization on the radial and mixed-inflow turbines for turbochargers J Zhang. A DOE table is created consisting of 25 different configurations. Initially there are 17 design parameters and 15 performance parameters. UK P Eynon Cummins Turbo Technologies. the design parameters are set up in such a way as to improve the efficiency while in design2 the aim is to reduce the turbine inertia. M Zangeneh Department of Mechanical Engineering. The resulting response surface is then used to investigate the sensitivity of different design parameters on the performance parameters. In design1. CFD and FEA is run on both designs and the results confirm that the design1 has higher efficiency than the baseline impeller but at expense of higher stress and lower impeller stiffness. UK ABSTRACT A methodology for designing radial and mixed-inflow turbines to meet multiple aerodynamic and mechanical requirements is presented in this paper. 2014 399 . By using the sensitivity information two different impeller geometries are designed. The method couples a 3D inverse design code and Design of Experiment Method (DoE) and the response surface method (RSM) to design turbines which meet various design criteria. Pierret [6] and Pierret et al. Furthermore. such as blade angles. For example. This creates a difficult multidisciplinary/multipoint problem that is rather challenging to meet by using direct design approach and trial and error iteration. In such an approach the rotor geometry is parameterized by using blade shape related parameters.1 INTRODUCTION The design of radial turbines for turbocharger applications poses difficult multi- disciplinary challenges. 2 OPTIMIZATION METHODOLOGY The flowchart of the optimization method that will be developed as part of the current project is shown in Fig. inverse design and RSM in the multidisciplinary optimization of turbocharger turbines. Such an approach can be coupled with Design of Experiments method and appropriate surrogate models such Response Surface Model or Kriging. Such approaches have also been used in compressor design. One possibility is to couple automatic optimization with direct design approach so that the design space is explored more systematically.[3]). these surrogate models will not provide accurate response surfaces. An alternative approach is to use an inverse design based optimization strategy as used for compressors in [8]. 400 . see [9] for further details. The advantage of such an approach is that a large design space can be explored with a small number of design parameters. One major advantage of the inverse designed based approach is that all the blade geometries generated in the design matrix will satisfy the specified turbine specific work at the specified mass flow rate. have low inertia and yet withstand high stresses at relatively high temperatures and have high stiffness against vibration. Even using surrogate models such as response surface or neural networks will require a large initial population. when large number of design parameters are used. to be used for optimization. [7] used a GA and RBF network for shape optimization of 3D compressor blades. Such a large number of design parameters will then require a large population size and hence many thousands of multipoint computations if an evolutionary optimization method is used. This will make it easier to achieve more accurate response surface. which can then be used for rapid evaluation of multi- disciplinary/multi-objective performance parameters required for optimization. Verstrate & Van den Brasembussche [5] combined a genetic algorithm (GA) and ANN to find a compromise between the conflicting demands of high efficiency and low centrifugal stresses for micro gas turbine blades. The turbine has to maximize power output from a pulsating flow from engine with variations in total pressure and temperature at inlet. The aim of the present paper is to present a design method integrated with DOE (design of experiment). Lian and Liou [4] developed a multi-objective optimization approach using second- order polynomial response surface model and evolutionary algorithm (EA) which was used in the redesign of a turbopump and the NASA rotor67 blade. Typically between 24-59 design parameters are required to represent the impeller (see [1] . 1. something that is difficult to control with direct design optimization. CFD and FEA calculations are performed for these designs to get the performance parameters yj. = 1. 2. = 1. by analysing and comparing the coefficients ai. we can eliminate non-essential design parameters compared to essential parameters and reduce the number of design parameters n. The performance parameters yj can be expressed by function of design parameters xi: = . 401 . In this paper. The performance parameters are used to evaluate the blade aerodynamic and mechanical performance which will be covered in the CFD and structural analysis section. 2. (2) The polynomial coefficients ai can be determined by a standard least-square regression. Figure 1 Flowchart of the optimization methodology The design and performance parameters are selected and denoted by xi and yj respectively. A linear DOE with at least (n+1) designs is needed to provide enough sampling points to solve the ai. … (1) Where n and m are the total number of design and performance parameters. the design parameters are parameters used to define the blade geometry which will be discussed in the blade parameterisation section. … . The first order polynomial RSM approximates the objective parameters yj with first order polynomial: = +∑ (2) There are total (n+1) unknown coefficients in the eq. Once the linear RSM model is available. Stacking condition. Blade loading distribution d(rVθ)/dm. 2. This method divides the loading distribution into three separate parts by NC and ND. 3 three segment method used to define the blade loading in TURBOdesign1 is illustrated. TURBOdesign1 (see [11]) is a three-dimensional inviscid inverse design method. The input design parameters required by the program are the following: Meridional channel shape in terms of hub. The baseline geometry is a radial turbine impeller for heavy duty diesel application with eleven blades. Fluid properties and design specifications. In Fig. blade tip diameter and blade length are constant in the optimization process. The blade has radial filaments as shown in Fig. It is imposed at two or more span locations. The code then automatically interpolates span-wise to obtain the two-dimensional distribution over the meridional channel. leading and trailing edge contours. Inlet flow conditions in terms of spanwise distributions of total temperature and velocity components.3 BLADE PARAMETERISATION The commercial software TURBOdesign1 [10] is used to parametrically describe the blade geometry. The stacking condition must be imposed at a chord- wise location between leading and trailing edge. the work coefficient is controlled. where the distribution of the circumferentially averaged swirl velocity (rVθ) is prescribed on the meridional channel of the blade and the corresponding blade shape is computed iteratively. Figure 2 Baseline Impeller Wrap Angle Contour The blade loading is defined as d(rVθ)dm. Everywhere else the blade is free to adjust itself according to the loading specifications. Parabolas are used to define the distribution for the first and last segment and the middle segment is 402 . the number of blades. shroud. Exit rVθ spanwise distribution. The circulation distribution is specified by imposing the spanwise rVθ distribution at leading and trailing edge and the meridional derivative of the circulation d(rVθ)/dm (blade loading) inside the blade channel. Due to external constrains (dimensions of nozzle and volute). Normal/tangential thickness distribution. By controlling its value. which is outlined in Fig. based on 6 parameters shown in red. At leading edge the spanwise rVθis kept fixed but trailing edge the spanwise rVθis modified. These parameters are inlet and exit width W1 and W2. In addition to blade loading the spanwise work distribution or rVθdistribution can be modified. Figure 3 Three segment method used to define the blade loading Figure 4 Meridional plane parameters Fig. In TURBOdesign1 a methodology is used for meridional shape parameterization. This adds additional two parameters to the optimization. Typically by defining blade loading in this way at two sections (hub and shroud) a very large variation in impeller shape can be obtained. Hence. Changes to Slope parameter can change the loading from fore-loaded to aft-loaded. 4 design parameters is all that is needed to represent the variation of in blade shape for a fixed meridional shape. 4. However. Using this parameterization. in this study an initial linear DOE will be performed and hence variations in NC and ND will also be included in the optimization. 5 shows possible configurations and variation of meridional plane in our design space. while changes to DRVT value at the leading edge will change the blade incidence. 403 . one can fix the value of NC and ND and still obtain large variation in blade shape by changing the Slope and the DRVT parameters.specified by a straight line with a given slope. Impeller leading edge and trailing edge angles and and two control parameters Y_hub and Y_shr that affect the meridional shape through the cubic spline method used to create the meridional shape. For saving computational time only one flow passage is modelled in the CFD which consists of a nozzle and a rotor as shown in Fig. MF_90k) for five different RPM are used to evaluate the turbine aerodynamics performance and flow capacity.5 Slope_Shr -1 . 404 . the shroud thickness remains same and the hub thickness is multiplied by the factor. The nozzle domain is stationary and rotor domain is rotating with a RPM ranging from 50.9 and 1. The total number of element is about 1.-5 DRVT_LE_Hub -1 .2 – 0.85 Slope_Hub 1 – 2. 6.000 rpm.05 – 0.04 RVT_TE_Shr 0.000.-0.06 – 0.9 – 1.05 – 0. So in total for the initial design table 17 design parameters were used and their variation ranges are shown in Table 1. T-S Efficiency (Eff_50k. The thickness between hub and shroud is interpolated proportionally.4 ND_Hub 0. The inlet boundary condition are total pressure and total temperature. MF_80k.5. Eff70k.11 (mm) W2 15 – 24 (mm) 0 . Figure 5 Meridional plane variation of the design space The blade thickness is controlled by one parameter called thickness factor which is between 0.1 NC_Hub 0. Base on the baseline thickness distribution.6 – 0.3. Table 1 Variation ranges of design parameters Design parameter Range W1 7 .000 to 90. Eff_80k and Eff_90k) and turbine mass flow parameter (MF_50k.85 ND_Shr 0.40 (°) 0 .2 NC_Shr 0.6 – 0.4 RVT_TE_Hub 0 – 0.10 (°) Y_Hub 16. MF_60k.000.1 Thickness 0. MF_70k.1 DRVT_LE_Shr -1 .-0. Eff_60k.5 – 21 (mm) Y_Shr 0. The flow direction at the inlet is specified by xyz components. Periodic boundary conditions are also applied. Atmospheric pressure is applied in the outlet.3 4 CFD ANALYSIS When the blade geometry is ready the CFD calculations are performed by using commercial software ANSYS CFX-14. Turbulence model is set as k-e. the mesh in the hub fillet is refined as shown in the Fig. To get more accurate results. The first mode shape for vibration is always exducer flex (left in the Fig. 8). 8) and the second mode shape is inducer and exducer flex (right in the Fig. 7.5 is used to perform structural analysis and modal analysis to get the maximum principal stress and natural frequencies. The maximum principle stress and 1st and 2nd order frequencies are chosen as performance parameters. It is found that the stress concentration always occurs near the hub fillet part (Fig. Therefore. whole wheel geometry is generated by ProEngineer (a popular CAD software) based on one blade. Figure 7 Mesh used in the structural analysis and stress evaluation 405 . Figure 6 Computational domain 5 STRUCTURAL ANALYSIS ANSYS Workbench 14. 7) and increasing of the fillet radius will reduce the local stress concentration. 6 to performed a linear polynomial regression. 25 designs in total are generated by DOE component in the ISIGHT5. =∑ + (3) Where is the ith design parameter (normalised to 0-1) and is the jth performance parameter (normalised to 0-1). CFD and FEA analysis are performed for these 19 designs. The range of the specific work for these 19 designs (RPM = 70k) is 57. 6 LINEAR DOE AND SENSITIVITY STUDY As shown in the section 3 we have in total 17 design parameters.3 – 66. Finally there are 15 performance parameters. 406 . moment of inertia and throat area which are other two performance parameters can be easily obtained from ANSYS Workbench. Table 2 shows the R2 value of linear RSM model for different performance parameters. 19 of 25 designs are converged and generate blade geometry by using inverse design code TURBOdesign1. The specific work is defined as turbine power divided by mass flow rate which can be obtained from CFD results. All the design parameter and performance data are imported into ISIGHT5.6 (Latin Hypercube method) based on specified design parameter ranges. In case some designs may fail to generate reasonable geometry. It can be seen that the error between the actual value and predicted value through the RSM model is very small due to the high R2 values (> 96%).85 kJ/kg versus baseline value of 65. Figure 8 Modal analysis At the same time.1 kJ/kg. To construct a linear (first order polynomial) response surface at least 18 different designs are necessary. 4 increasing of W1 will result in increase of nozzle inlet area and more flow can enter the nozzle. MF_60k. the coefficients is divided by the max | | and multiplied by 100 to get Ai (as shown in Table 3). MF_70k. = × 100 (4) (| |) Table 3 Comparison of normalised coefficients Aji The range of Ai is from -100 to 100. Table 2 R2 of the linear RSM model R2 Eff_50k 0. Therefore.9989 MF_70k 0.9685 MF_50k 0. the more significant the corresponding design parameter is.9984 Throat_area 0. When Ai is less than 0.9895 Eff_60k 0.9971 MF_90k 0.9977 1st_frequency 0.9896 Eff_70k 0.9991 2nd_frequency 0.9981 MF_80k 0.9749 Eff_90k 0.9802 Eff_80k 0.9973 Stress 0.9882 MOI 0. It can be easily understood by seeing the Fig.9863 To compare the effect of different design parameters (xi) on the performance parameter y. the flow capacity of the turbine will be 407 . MF_80k and MF_90k. the greater of the absolute value of Ai is. For the MF_50k. it means the performance parameter will decrease with the increasing of the design parameter.9991 MF_60k 0. When Ai is greater than 0. For a particular performance parameter. the most significant parameter is W1 since the values of Ai are 100. it means the performance parameter will increase with the increasing of this design parameter. design1 has 54% higher stress. 10. W1. The resulting meridional shapes of design1 and design2 are shown in Fig. the blade stiffness is reduced and 1st_frequency and 2nd_frequency will drop. W2 and Y_Hub. The values of other design parameters will be set same as the baseline value or intermediate value in the Table 1. This can also be explained by the Ai value of for Eff_80k and Eff_90k are 100. the total MOI will increase. Therefore. The results confirm that design1 results in 2-5 points improvement in efficiency as compared to the baseline impeller and also about 5-7% increase in flow capacity. Similarly if we want to minimize MOI we need to maximize and minimize W2. 2. Y_Hub in the design space.54% increase and 31% reduction in modal frequencies. The aims of design 1 and design 2 are to maximize Eff_70k and minimize MOI (moment of inertia) separately. But as shown in Table 4. Similarly for the MOI. Since the variation of these 7 design parameters will have a much larger effect on the performance parameter compared to the other design parameters. 9 and compared with the baseline impeller. increasing of W2 and Y_Hub will increase the blade inertia and wheel hub inertia. The results confirm clearly that any optimization of the turbine would require multi- objective optimization approach that can find the best tradeoff between these contrasting requirements. shown in blue. However. Y_Hub in the design space. Our CFD calculations show that with the increasing of mixed-inflow turbineU/C with peak efficiency moves towards higher RPM (80 or 90k). Figure 9 Comparison of Meridional Plane 408 . design2 has a significantly narrower exducer height. Design2 results in 30% reduction of inertias as compared to baseline but almost 15 % loss of efficiency versus the baseline design. By comparing Ai values in row Eff_70k from Table 3 if we want to maximize Eff_70k we need to maximize NC_Hub. NC_Hub. Design1 has a mixed flow inlet with wider inlet width and with almost the same exducer height as the baseline impeller.7 corresponding to RPM = 70k in our design problem. Slope_Hub. Two designs are generated for these 7 selected design parameters based on Table 3 to validate this method of parameters reduction. The comparison of the total- static efficiency and mass flow parameter (normalized by the baseline efficiency and flow parameters at 70k rpm and 50k rpm) is shown in Fig.improved. ND_Hub. Therefore. It is well-known that the peak efficiency of a radial turbine () occurs when the U/C (velocity ratio or jet speed ratio) = 0. The 7 most significant design parameters are selected through summation of all the absolute value of Ai which are α1. Work is currently underway to develop such an approach for radial and mixed flow turbines. if we increase the W2 which is the blade height at the outlet the blade will become higher and easier to vibrate. For the 1st_frequency. W2 and minimize ND_Hub. Kriging approximation will be used to construct a more accurate model based on the new 7 design parameters using a higher number of configurations. However. A DoE table consisting of 25 configurations is generated by using the 3D inverse design method and using 17 different design parameters that affect the meridional shape. The performance parameters for these two designs were obtained by CFD and FEA and generally they confirm that the sensitivity information provides the correct direction in terms of the design objectives. stress.8% 11. vibration and inertia. the blade loading. maximum stress and 1st an 2nd modal frequencies and impeller moment of inertia. 409 . The methodology couples a 3D inverse design method.4% -31.0% Design 2 -30.2% -38. By setting these 7 parameters appropriately two designs were created one for higher efficiency and one for lower inertia.54% 55.2% 7 CONCLUSION A detailed methodology about blade design and optimization has been shown in this paper. the spanwise work distribution and the thickness distribution. The results was then used in a sensitivity analysis using a linear response surface to find the 7 most significant design parameters. In future work. Figure 10 Comparison of T-S Efficiency and Mass Flow Rate (normalised) Table 4 Structural performance compared to the baseline value Moment of Inertia Stress 1st frequency 2nd frequency Design 1 2. A Pareto Front will be generated by applying MOGA on the Kriging model with multiple objectives and constraints.9% 33.4% -31. A full optimization of turbine would require a proper multi-objective optimization. the results also indicated that design1 that improves efficiency also results in increase in stress and reduction in modal frequencies. design of experiments and linear response surface model in order to find the most important design parameters that affect the efficiency and mass flow parameter at multiple operating points. By using this approach it should be possible to obtain the final optimized geometry that meets the contrasting requirements of aerodynamics. for multidisciplinary design of radial and mixed flow turbines for turbochargers. Each configuration was then analysed in CFD at multiple operating points and FEA for structural and modal analysis. Bohle. Proceedings of ASME Turbo Expo 2012. Multidisciplinary Optimization of a Turbocharger Radial Turbine.. Vol 135. Multiobjective Optimization Using Coupled Response Surface Model and Evolutionary Algorithm. A 3D Design Method for Radial and Mixed-Flow Turbomachinery Blades. On the Coupling of Inverse Design and Optimization Techniques for the Multiobjective. T. & Van den Braembussche. T. The Influence of Metamodeling Techniques on the Multidisciplinary Design Optimization of a Radial Compressor Impeller. 2009.. Multipoint Design of Turbomachinery Blades. Multidisciplinary Design Optimization of a Mixed Flow Turbine Wheel. [2] Chahine. GT2012-68233. L. 1991. Multidisciplinary Optimization of a Radial Compressor for Micro Gas Turbine Applications. 2013.Advanced Design Technology Ltd. The author would like to thank their financial support. [10] TURBOdesign1 . [5] Verstraete. ASME Turbo Expo 2007. Journal of Turbomachinery. 33: 61-70.. Filomeno Coelho.. 410 . & Kato. European Conference for Aerospace Sciences. R. 13: 599- 624. H. 2005. Optim. [7] Pierret. Vaidyanathan. T. Alsalihi. & Gugau. & Liou. & Verstraete. M. 2007. C. T. [3] Muller... Seume. J. R. Progress in Aerospace Sciences. ALsalihi. M. & Tucker. [4] Lian. 2012. R. Multi-objective Optimization of Three Dimensional Turbomachinery Blades. Struct. [11] Zangeneh. Multidisc. Journal of Turbomachinery. Z. M. & Verstraete. D. REFERENCES [1] Roclawski. 2007. & Zangeneh. Goel. ASME Turbo Expo. M. Haftka.. H. S. Multidisciplinary and Multiple Operating Points Shape Optimization of Three-dimensional Compressor Blades. A. P. Z. GT2012-68358. R.ACKNOWLEDGEMENTS This research is sponsored by Cummins Turbo Technologies and EPSRC. M. [9] Queipo. Vol 131. 2005. N. S.. 43(6): 1316-1325. Shyy. K. 41: 1-28. 2013. [8] Bonaiuti. R. T.. 2005. Y. W. AIAA Journal. 2012. V. Journal of Numerical Methods in Fluids. Surrogate-based analysis and optimization. [6] Pierret. S T Liu. This may be a concern of high-cycle fatigue failure of VNT rotor blades. As a result. Through FFT analysis of the pressure fluctuation at the rotor blade leading edge. Y H Liu.Numerical investigation on the high-cycle pressure fluctuation mechanism of VNT rotor X Shi. In this paper the 3-D unsteady numerical simulation of a VNT was performed to investigate the effects of nozzle vanes clearance leakage. 1 INTRODUCTION Turbocharger has been widely used in diesel engines due to it has opportunities to improve engine power and fuel economy as well as reduce emissions[1. The shock wave only affects the suction surface at the leading edge of rotor blade because of the shape of rotor blade. This may be a concern of high-cycle fatigue failure (HCF) of VNT rotor blades. a better method to meet this requirement is the use of a variable geometry turbine (VGT) which can increase pressure ratio and adjust flow angle at low mass flow condition by varying throat area of nozzle. Thus leakage. the magnitude of the pressure fluctuation is increased significantly by the influence of the nozzle vanes clearance leakage. stronger emphasis from emissions legislation was forcing automotive manufactures to improve part load performance of the engine.6]. In the regions near the hub and shroud side on the rotor blade suction surface. Beijing Institute of Technology. which can specify where and how these three factors influence the excitations of the turbine rotor blade. together with wake and shock wave. But the strong excitations at the leading edge of turbine rotor can be created during engine braking down or transient acceleration. the shock wave and the amplitude of the pressure fluctuations on the rotor blade surface were greatly increased compared to that for full open condition. When the nozzle vanes was closed to small openings and turbine operates at high pressure ratio. Leakage and wake from nozzle vanes have significant impact on flow distribution at inlet of downstream rotor. The results show that the shock wave and nozzle vanes clearance leakage are the main reasons for rotor blade surface pressure fluctuation. it can be seen that the fluctuation amplitude increased significantly on the rotor passing frequency and this also implies the possibility of VNT high-cycle fatigue failure with the high-frequency excitations.2]. turbine performance and unsteady load of rotor blade[3. B Zhao School of Mechanical Engineering. 2014 411 . wake and shock wave on the rotor blade pressure fluctuations. In the last two decades. the strong excitations at the leading edge of turbine rotor can be created during engine braking down or transient acceleration. When the nozzle vanes opening was closed small.4]. C Yang. contribute to _______________________________________ © The author(s) and/or their employer(s). China ABSTRACT Due to the rotor-stator interaction of Variable Nozzle Turbine (VNT). shock wave appears on the suction side of nozzle vane near trailing edge and induces high pressure response on the surface of blade by impinging at the leading edge[5. Jameson’s dual time step method is used.3 Boundary conditions and convergence Total pressure. and then specify where and how these three factors influence the excitations of the turbine rotor blade. flow is supposed to be fully turbulent and one equation Spalart-Allmaras model is used for closure of RANS equations.35 mm. Cell centered control volume approach is used for central space. Static pressure was specified at the rotor outlet boundary with radial equilibrium. Throat area of the nozzle vane can be changed by rotating the adjusted shaft. Steady simulations showed that the changed blade number causes a small deviation in turbine performance. The rotating speed of rotor is 120 krpm. For steady computation. Each of nine nozzle vanes equips with two clearances with 2% of spanwise height of vane. 3 NUMERICAL METHODOLOGY 3.0 is used in all the computations. After scaling the computational domain. total temperature as well as flow angle calculated according to volute are imposed at the nozzle inlet. this means that the nozzle vanes’ opening condition is various. In this paper a 3-D unsteady numerical simulation of a VNT was performed to investigate the effect of nozzle vanes clearance leakage. and other surfaces are stationary. wake and shock wave on the rotor blade pressure fluctuations.HCF problem of downstream rotor. A 4-stage Runge-Kutta integration scheme is used for time discretization in steady-state simulation. CFL number of 3. 3. The volute was not included in the computational domain because of current project researches the interaction between nozzle vanes and downstream rotor. Three dimensional time-averaged Navier-Stokes (RANS) equations were solved.2 Numerical methods The EURANUS flow solver integrated in NUMECA Fine/Turbo is employed in this work. As their features are different each other. 1. momentum and energy through the interface. They locate respectively on the hub and shroud sides. Differential scheme and Jameson type of artificial dissipation items are used for scheme stability. the conservative coupling by pitchwise rows was used at the rotor/stator interface.01 are taken separately for 2nd and 4th order items. 120 physical time steps are set for the computational domain. 2 DESCRIPTION OF TURBINE The variable geometry turbine (VGT) used in this investigation is originated from a turbocharger with mixed rotor.39 million cells are used for the numerical simulation. A method to applied domain scaling method is to change number of rotor blades from 13 to 12 to meet the required condition that both sides of rotor-stator interface are same periodicity. Totally 2. The computational grid was generated automatically by IGG/Autogrid5. In current case. For the unsteady computations. This approach adopts the same 412 . only 1/3 of blade rows consisted of three nozzle vanes and four rotor blades was modeled in this paper. Rotor blade’s tip gap locates only on the shroud side and is about 0. For the artificial coefficients. which guaranteed an exact conservation of mass flow. Adiabatic condition is set for all the solid walls.0 and 0. 3. Thus there are 9 nozzle vanes and 13 rotor blades in this researching turbine stage. it is necessary to know in detail the progress that each of them induces highly aerodynamic loading on turbine wheels.1 Computational grid Domain scaling method is used for geometry configuration and computational mesh to reduce requirement of CPU time and space of memory. Only downstream rotor blade surface and hub rotate. however. From these two pictures it can be seen: for large opening only one shock wave stands on the suction side of nozzle vane while two shock waves appear in small opening.05 0.0 Pressure ratio 0. resulting in that predicted efficiency is about 2.6 0.0 0. Figure 2 Distribution of static pressure at mid-span 413 .03 0.01 0.5 percent higher than tested data.5 0.60 Efficiency 3.45 1.0 Cal Cal Exp 4. Compared with large opening. 3.04 0. one locates at throat of nozzle and another one stands on suction surface of nozzle vanes near trailing edge. For the unsteady computation.5 0.coupling procedure for all the nodes along the circumferential direction. a direct interpolation on sliding meshes was employed at the rotor-stator interface.70 5.5 present higher than the test pressure ratio.5 2.06 U/C Dimensionless flow rate Figure 1 Comparison of predicted turbine performance with test result 4 ANALYSIS AND DISCUSSION 4.02 0. The volute was not included in the simulation model. As this.65 4. Figure 1 show the comparisons of turbine performance which is from CFD and experiments respectively.1 Comparison of shock wave for two kinds of opening Figure 2 shows the distribution of static pressure with isoclines at mid-span for two kinds of open conditions. Notice that shock wave can impinge at the leading edge of downstream rotor blade for any opening.8 0.5 Exp 0. 0.7 0.5 0. The predicted mass flow of turbine is about 3.4 0. This is mainly due to the difference between computational domain and test hardware. the flow loss in volute was not calculated. current CFD method which is able to predict well turbine performance can be applied to assist turbine analysis.0 0. Thus stronger interaction between shock wave and blade occurs when nozzle vanes are adjusted to small opening. even if the local flow direction is different from the average one. the shock wave stood on the suction side of nozzle vane is much stronger in small open condition.4 Verification on numerical results Only steady CFD method was used for validation by comparing the numerical results with experimental data. Despite of this discrepancy.55 3.50 2.40 0. at small opening 4. the inlet of rotor blade is radial and the shock wave can only impinge a small area near the leading edge of suction surface. Wake generated by nozzle vanes flows circumferentially and experiences long distance before entering into downstream rotor channel. The leakage rolls up and then forms vortex. the pressure distribution on the blade surface must be changed. the flow between nozzle vanes and downstream rotor is highly unsteady. during downstream transportation. meanwhile. t1 t2 t3 t4 t5 Figure 3 Interaction between shock wave and rotor. the leakage flow from nozzle vane not only changes the flow angle entering rotor. C and D in Figure 4.2 Nozzle vanes’ leakage The nozzle vane’s leakage flow increases as the nozzle opening is closed small. Another thing should be noticed is different with axial turbine. Downstream part transports fast along suction surface of rotor blade. The shock wave generates when the smallest width between nozzle vane and rotor forms. So only small opening in this paper is used for the analysis of the interaction between nozzle vane’s leakage and downstream rotor. such as the region marked with A. At this time. Another part interacts with boundary layer on the pressure side of rotor blade. as indicated at time of t1 (t1~t5 correspond to different instantaneous time). The air flows from vane’s pressure side (convex surface) to suction side (concave surface) and then forms leakage in the downstream region of nozzle vane. the leakage can also influence downstream rotor like shock wave by the mean of hitting at leading edge. As downstream rotor blade passes. so the width of region with high entropy gradually increases. This long track wake transports 414 . It should be noticed that on the pressure side of blade transportation speed of leakage is not as high as that on suction side.When the turbine operates in small opening. When the leakage hits at the leading edge of downstream rotor. the strong interaction between shock wave and leading edge of rotor is easy to induce high cycle fatigue (HCF) of downstream rotor because of highly unsteady pressure response must have been induced when shock wave impinges at the blade. the shock wave begins to weaken. but influences the development of boundary layer of blade as well. It has little effect on pressure side of downstream rotor blade. In addition to mixture with surrounding flow. resulting in higher entropy appears in the center of leakage flow. the time nozzle vane’s leakage interacts with pressure side of blade is longer. This progress has been shown well by several patterns signed with B. the intensity of shock wave changes. As the leading edge of rotor blade leaves away. as indicated in Figure 3. These features have been shown well by distribution of entropy in Figure 4. the leakage continuously mixes with main flow and leads more flow loss. Moreover. As known. more flow loss was induced on the suction side of rotor blade. This progress is clearly shown in Figure 3. This fact has been indicated by many researchers[3. Figure 5 shows the distribution of entropy at midspan for small opening. Meanwhile. Besides.7-9]. possibly leading to strong pressure response in this region. This is why nozzle vane’s leakage influences HCF of downstream rotor. the leakage was cut into two parts. the strength of shock wave increases nearly to maximum and the rotor blade suction surface near leading edge experiences serious interaction from shock wave. As a result. But. at 90% span Figure 5 Distribution of entropy. it has been known that the shock wave.3 Aerodynamic responses on downstream rotor From above analysis. Even though blade force is not the only parameter affecting the force response [10]. the small space must enhance the aerodynamic interaction and then increase risk of HCF of rotor.provides more time to mixture between wake and surrounding air. Figure 4 Interaction of nozzle vane’s leakage and rotor. interact with downstream rotor when turbine operates at small opening. Moreover. those strongly aerodynamic interactions still influence seriously HCF of 415 . Thus this contradiction should be taken into account in design progress. the more flow loss mixture leads. the small space between nozzle vanes and downstream rotor should be available to improve the turbine performance. Thus wake is weak when it hits at leading edge of blade. According to this phenomenon. the longer distance wake transports. and its influence on the pressure response on the surface of blade is not as serious as shock wave and leakage flow. on other hand. at midspan 4. together with nozzle vane’s leakage. small opening 4. This is due to the fact that shock wave and leakage mainly contact the leading part of blade. The larger RMS value is. From these two pictures can be seen that the regions with high RMS mainly concentrate in the inlet of blade. This must increases the risk of blade fatigue. Figure 6 shows the root mean square (RMS) of pressure fluctuation on both pressure and suction sides of the rotor blade for small opening. This fact implies that VPF is most dominant frequency among several excited frequencies if VNT suffers from high-cycle fatigue failure due to the high-frequency excitations. Harmonic analysis was carried out to calculate the pressure response at several monitoring points on the blade surface. which is three times of the VPF. (a) pressure surface (b) suction surface Figure 6 RMS pressure distribution on the blade surface. From this result it can be known that dominant frequency is found at frequency of 18 kHz. Based on this fact the Fast Fourier Transformation (FFT) method has an advantage to reflect pressure response at any one frequency. its number can be adjusted in a large scale while few selections about the vane number can be provided for thick vane. strong shock wave appears and impinges at the leading edge of blade. Subordinate signal is found at frequency of 36 kHz. the stronger pressure fluctuation is. This is the reason that only one kind of nozzle vane number was modeled and commutated in current article. due to nozzle vane equips with two gaps which locate respectively on hub and shroud sides. It is also noted that the VPF is associated with vane number. the pressure in the regions near leading edge is more sensitive to nozzle vane’s influences.4 Distribution of pressure fluctuation RMS of pressure is able to reflect information about pressure fluctuation on the blade surface. For the thin nozzle vane.[11] had applied numerical methods to examine the aerodynamic excitation of the blades in radial turbines with guide vanes. In other word. Bauer et al. which is the same as nozzle vane passing frequency (VPF). For the suction side RMS pressure distribution. with pressure pulsation amplitude being about 570 kPa. with pressure pulsation amplitude only being about 140 kPa. but it includes all information that pressure fluctuates at many frequencies. Pressure spectrum result on the leading edge of rotor blade at midspan is shown in Figure 7. As analysis above. Compared with pressure side. which is twice of the VPF. and the results revealed that unsteady loads appear significantly higher for 12 compared to 22 vanes. the leakage flow from both gaps leads to two regions with high RMS near hub and shroud. 416 . Other signal with smaller pulsation amplitude can also be found at frequency of 54 kHz. It results in large region with high RMS appears near midspan location. stronger pressure fluctuation appears on the suction side near leading edge.blade by mean of inducing pressure response on blade surface. This means that nozzle’s influence on the downstream rotor weakens gradually along the flow path. mid span Through FFT analysis of the pressure fluctuation at six predicted points on the both surfaces of rotor blade at midspan. Illustration of monitoring points 600 600 500 500 Amplitude /kPa Amplitude /kPa 400 400 300 300 200 200 100 100 0 0 1 2 3 4 5 6 1 2 3 4 5 6 Location of monitoring points location of monitoring points Pressure side Suction side Figure 8 Pressure pulsation magnitudes at nozzle VPF at monitoring points 5 CONCLUSIONS Shock wave and nozzle vanes clearances’ leakage are the dominant reasons for pressure response on rotor blade surface. A fact along meridian direction the magnitude decreases gradually can be seen. Figure 7 Pressure spectrum at rotor leading edge. This fact also indicates that nozzle vanes have more influence on the suction surface of rotor blade. In addition. the amplitudes of pressure pulsation at VPF on the suction side are far larger than that on pressure side. the amplitudes of pressure pulsation at frequency of VPF were obtained and shown in Figure 8. and the influence from nozzle vane’s wake is not as strong as shock wave and leakage because of it mixes fully with surrounding 417 . Proc.. 24. Of ASME Turbo Expo. “An investigation of the flow field through a variable geometry turbine stator with vane endwall clearance”. (3) Hu. G. 1982. N. GT2008-50461. Neill. Sun. Wiley Interscience. In the regions near the hub and shroud side. L... “Turbocharging the internal combustion engine”. 2012. Unno. 4: pp. (7) Walkingshaw. 2012. “An experimental assessment of the effects of stator vane clearance location on an automotive turbocharger turbine”. (9) Tamaki.C.. When the nozzle vanes locate at small open condition. (2) Baines. Thornhill. C. et al. 2011. N. et al. (5) Chen. 10th international conference on turbochargers and turbocharging. 220: pp. This article is the first part of our researching project. Proceedings of the Institution of Mechanical Engineers Part A: Journal of Power and Energy. Yi. Of ASME Turbo Expo. 1982. Of ASME Turbo Expo. H. Institute of Mechanical Engineers... it can be known that the dominant frequency of pressure pulsation is of same as the nozzle vane passing frequency.... S. et al. et al. 2006.. Vol. MTZ worldwide. New York. Institute of Mechanical Engineers. 6: pp. 2006. 2013.. 2013. “Turbine wheel design for Garrett advanced variable geometry turbines for commercial vehicle applications”. vol. 2013-01-0918. The CFD/FEA coupled forced response analysis had been done. GT2013-94761. 2006. “Investigation of Nozzle Clearance Effects on a Radial Turbine: Aerodynamic Performance and Forced Response”. REFERENCES (1) Watson. 8th international conference on turbochargers and turbocharging.. GT2010-23677. 2013 SAE International. but it was not included in this paper due to the limited length of paper.. Chinese Journal of Mechanical Engineering. “Radial and mixed flow turbines options for high boost turbocharger” 7th International Conference on Turbocharger and Turbocharging. Cunningham. Yang... H. 899-910. J. Moreover. 2006. 2002. “Prediction of high-frequency blade vibration amplitudes in a radial inflow turbine with nozzle guide vanes”. T. 51276018). Through FFT analysis of the pressure fluctuation on the rotor blade surface..S. “Numerical Analysis of Nozzle Clearance’s Effect on Turbine Performance”. S. Janota. Proc. “The effect of clearance flow of variable nozzle on radial turbine performance”. 74. (10) Schwitzke M. Sun. (8) Spence. Harold Sun of Ford for his strong technical support. 618-625... J.. Goto.. Schulz A. “Unsteady rotor-stator interaction of a radial inflow turbine with variable nozzle vanes”. Bauer H. the shock wave increases significantly compared with large open condition.. S....W.. Spence. London UK. (6) Kawakubo. the pressure fluctuation at leading edge responded to leakage from nozzle vane’s clearance is strong at small opening. Schwitzke M. vol. shock wave mainly influences the pressure response on suction surface of blade near leading edge. R.W. 2010. H.flow during transportation in the space between two rows. (4) Hu. D. Proc. H. Schulz A. M. (11) Bauer H. 418 . L.. “Aerodynamic excitation of blade vibrations in radial turbines”. ACKNOWLEDGMENT This research was supported by Ford Motor Company and National Natural Science Foundation of China (NSFC. 2011. W. J. M. London UK. The authors would like to acknowledge their financial support and Dr. 48-54. ωS ring and shaft angular speed P film pressure field Ω ring speed ratio (Ω = ωR/ωS) 1 INTRODUCTION Today. reliable and very cost-effective due to its simplicity and _______________________________________ © The author(s) and/or their employer(s). A complex thermo-hydrodynamic simulation model is introduced which is validated by experimental data under real TC operating conditions on a hot-gas test stand. improve the engine efficiency and keep the emissions below the legal regulatory limits. Do ring inner / outer diameter Tsup TC lubricant supply temperature F film fill factor Tshaft shaft temperature boundary condition h film thickness η mean effective lubricant viscosity nshaft TC shaft speed ρoil lubricant density psup TC lubricant supply pressure τD drag torque ∆p pressure loss φ circumferential coordinate (angle) pl. Technical University of Clausthal. The rotors of turbochargers used in passenger vehicles are typically supported by two journal bearings of the floating ring type which consist of a short cylindrical bushing placed between the shaft and the turbocharger housing forming two thin fluid films in series (Figure 1). The conducted study clarifies the significant impact of the shaft temperature on overall bearing operating parameters which cannot be neglected. This bearing system lubricated with internal combustion engine oil has proven to be robust. Ro ring inner / outer radius Di. turbocharging the internal combustion engines of passenger cars is one of the most promising technologies to increase the power output.Thermal analysis of small high-speed floating-ring journal bearings D Porzig1. 2014 421 . Bo ring inner / outer width Ri. Germany ABSTRACT In the present paper. J R Seume2 1 Institute of Tribology and Energy Conversion Machinery. H Schwarze1. H Raetz2. Germany 2 Institute of Turbomachinery and Fluid Dynamics. the impact of thermal boundary conditions on characteristic parameters of full floating-ring journal bearings (FRB) typically used in small automotive turbochargers (TC) is discussed.i inner film supply pressure ωR. the accuracy of ring-speed prediction is considerably improved by a non-adiabatic approach. Leibniz Universität Hannover. NOTATION A sliding surface area Q lubricant flow rate Bi. Compared to using an adiabatic shaft model. Integrated with an internal combustion engine. This results in non-uniform temperature distributions in the shaft. San Andrés et al. Bohn et al. The model 422 . 8. In automotive turbochargers. and the bearing system itself which is complicated to analyze because of the multiplicity of different heat flow paths [3. 4]. 9] used a 3D Conjugate Heat Transfer (CHT) model with experimental validation to advance the understanding of the heat balance in automotive turbochargers but without direct coupling to a bearing friction model including heat flux. Ring speeds predicted by isothermal flow models exhibit poor correlation with respect to test data especially at high shaft speeds [12]. The models used to characterize the thermal state of the TC were continuously refined in later publications. where the rotation of the ring is inhibited by an axial pin. a full-floating bearing has the advantage of reducing the rate of shear in the fluid films. 11] developed a model incorporating a lumped-parameter thermal energy balance for the estimation of the lubricant viscosity and change of operating clearances in the fluid films. thus reducing the total frictional power loss [2]. Trippett and Li [13] presented a thermal model indicating that thermal effects significantly affect the performance of high-speed floating-ring bearings and cannot be neglected. [7. Current automotive TC performance prediction tools include diabatic TC models [5. the bearing system is exposed to a difficult thermal environment [1]. 6] using lumped-parameter 1D approaches for the thermal energy flows to improve TC operating predictions mainly in part load and transient simulations. In comparison with a semi-floating ring bearing (SFRB). Figure 1 . Besides the shear-driven frictional losses within the fluid films acting as heat sources. [10.Sectional view of a typical turbocharger for passenger vehicles and schematic illustration of the full-floating ring bearing system Early investigations of (semi) floating ring bearings (summarized in [10. the turbine wheel is also exposed to hot exhaust gas from the engine which reaches temperatures of 600°C in modern Diesel engines and higher temperatures in modern Otto-cycle engines. the TC casing. 11]) were limited to experiments at low shaft speeds used for verification of isothermal analyses only. the turbocharger’s bearing friction plays an important role for overall system efficiency and behavior because friction leads to an inertial lag upon acceleration and produces backpressure on the engine caused by the turbine wheel [1].ease of manufacturing [1]. As a second important aspect. The temperature distribution in the fluid films. The most recent thermohydrodynamic model for (semi) floating ring bearings presented by San Andres et al. The model predicts pressure. the shaft is a source of thermal energy which was assumed to be adiabatic due to missing reliable test data. The thermal model presented does not include non-uniform temperatures in the shaft and the TC casing but it assumes uniform casing and shaft surface temperatures over the whole operating range. The temperature distribution in the floating ring is taken as uniform in the circumferential and axial directions and considers heat flow in radial direction only. a sophisticated experimental and numerical approach for the determination of thermal boundary conditions is described. 2 (left) for reference. steady-state model is introduced in the present paper. Boundary conditions for the solution of the Reynolds equation are ambient pressure at the bearing discharge planes and supply pressure at the individual oil supply holes in the outer and inner films. The authors pointed out that in an actual TC driven by hot gas. An iterative algorithm finds the balance between the external radial static load on the 423 . The model is used to predict the performance of a typical semi floating ring bearing for a TC application with shaft speeds of up to 240 krpm.predicts the bearing performance based on an effective lubricant viscosity and assumes constant film and ring temperature throughout the flow domains. no comparison with measurements from a hot gas test was shown. A mass conserving cavitation model based on the algorithm by Elrod [16] locates the cavitation boundaries and gives information about the local oil film fill factor throughout cavitating lubricant films. Also. revealing a large radial temperature gradient across the floating ring. The resulting inner and outer film pressure distributions are shown in Fig. The pressure distributions in the outer and inner fluid film are determined by solving a two-dimensional (2D) extended and generalized Reynolds equation [14. The importance of precise thermal boundary conditions for accurate bearing performance predictions is shown in detail below. temperature and viscosity distributions in both fluid films and analyzes heat flow paths according to thermal boundary conditions. For a systematic investigation of thermal boundary conditions and their impact on FRB performance parameters. floating ring. [3] models the thermal energy transport in a (semi) floating ring bearing system and predicts pressure and temperature distributions in both fluid films. While the model points out the most relevant thermal effects on FRBs and gives a good overview of the heat flow paths. 15] considering the effective hydrodynamic angular velocity for each individual film. 2 THEORETICAL ANALYSIS The floating ring bearing system performance parameters are calculated by a numerical simulation program. and heat flow distributions. drag power losses. The comparison of predicted ring speed ratios with measured values from a cold gas test rig showed a fair correlation but underestimated the decrease of the ring speed ratio Ω for high shaft speeds. a complex thermo-hydrodynamic. shaft and an annular TC casing segment is calculated three- dimensionally which provides for an in-depth investigation of thermal effects like varied boundary conditions and their influences on bearing performance parameters. The predicted performance parameters for varied thermal boundary conditions and heat flow concepts are later compared with actual measured test data from a hot gas test. the predictions are made for a semi floating ring bearing system only. Integrating the individual pressure field over the corresponding bearing surface returns the fluid film reaction forces resulting from the hydrodynamic pressure and hydrostatic oil supply [3]. the radial ring temperatures and other performance parameters like operating film clearances. respectively. Due to high rotational speeds. incompressible and laminar oil film. a two-dimensional energy transport equation is sufficient for proper modeling of the heat flow in the shaft segment [14]. 10]. a continuous heat flux is postulated. explained in detail in reference [14]. The maximum film temperature develops towards sides of the bearing in both films. This makes the calculation of a 3D temperature distribution in both semi-floating and full-floating rings necessary for precisely modeling the heat transfer through the ring. Fig. the influence of the oil feed holes to the inner film will produce local spots of lower temperature along the ring circumference as well.FRB and the fluid film reaction forces by iteratively modifying the displacements of journal and floating ring. Estimation of the effective clearances is performed by considering this thermal expansion of shaft. exactly modeling heat transfer between lubricant and solid bearing surfaces and neither requiring calculation of local heat convection coefficients nor iterative balancing of heat flows between bearing sections. 1 (right) for reference). Unlike most previous thermal FRB models [3. In full-floating rings. The local film fill factors are used to scale the heat transfer at the interfaces to account for the lowered heat transfer capability in partly filled film regions [16]. Operating film clearances in the inner and outer film contact change during operation due to thermal expansion of the individual bearing solids (shaft. the solid surface temperature and the clearances at room temperature. the present computational program also solves 3D energy transport equations for an annular segment of the casing and the ring (model scope shown in Fig. Semi floating rings in particular show a significant temperature gradient in circumferential and axial directions due to non-uniform temperature distributions in the adjacent fluid films. Since the Reynolds equation is solved two- dimensionally only. The outlined approach presents an efficient and precise method to calculate the temperature distribution in the FRB. the temperature is set to be circumferentially uniform at the outer ring and shaft surfaces. 424 . 2 (right) depicts the resulting temperature distributions at the inner and outer film centers for a symmetrical FRB (cut at the axial mirror plane). the temperature is significantly lowered. At the interfaces between the single sections of the bearing. ring and TC casing). In those regions of the film where fresh lubricant enters. taking the individual thermal material properties into account. The TC casing will usually develop a temperature variation across the circumference as the wide oil supply inlet acts as a local heat sink while conductive heat transfer through the TC casing from the turbine heats up the surroundings of the bearing unevenly as is confirmed by test data below. Continuity of the temperature distribution is preserved at the interface by setting the local lubricant temperature equal to the temperature of the solid surface. The 3D temperature distributions in the lubricant films and the associated Reynolds equations are coupled through temperature-dependent. local effective lubricant viscosities that vary in all three dimensions. The clearances are calculated using a simple model also shown in [3] and are updated on every completed solution of the temperature distribution in the bearing system. The temperature distribution in the fluid films is derived by solving a full 3D energy equation for a steady. As the shaft temperature does not vary in the circumferential direction due to the high rotational speed. the radial variance of the effective lubricant viscosity is therein considered using correction factors [14]. ring and casing which depends on the specific material characteristics. The movement of the holes presents a disturbance of the hydrodynamic pressure in the outer lubricant film as the oil flow into the inner film leads to local collapse of the outer film pressure. hi) and fill factor (Fo. Pi). Simplified equations for drag torque calculation for both film flow regions are given in (1) and (2). nshaft = 80 krpm) Determination of the floating ring rotational speed is based on a balance of the shear-driven torques τD. o ho Po D. The local fill factor equals 1 for fully filled film regions and has a value between 0 and 1 in areas where the film cavitates.i on the inner and outer ring sliding surfaces [10].o and τD. psup = 3 bar(abs). The drag torques are a function of the local film parameters pressure (Po.i 425 . proportional to the rotational speed of the ring results from centrifugal forces [17]: oil pc 8 R2 Do2 Di2 (3) As the half-moon grooved supply plenum spans the TC casing surface only partially. A hydraulic film interfacing model for FRBs is used to transform the transient processes involved into a steady-state model. The field of constant pressure creates a drainage area allowing a volumetric oil flow Qo. not all of the feed holes will face the outer supply groove during ring rotation which leads to a pressure deviation between the feed holes. mean lubricant viscosity (ηo. scaling the local shear to take the lowered friction in partly filled film regions into account [16]. The fluid films are hydraulically interfaced through the radial lubricant feed holes in the floating ring circumference. In the simulation program.o Ro F Ao o ho R Ro Fo dA 2 o (1) i hi Pi D.Figure 2 .Pressure distribution (left) and temperature fields at film center (right. the circumferential area where the radial feed holes traverse is set to a uniform pressure pl. a pressure drop across the feed holes ∆pc. only half bearing shown) for both lubricant films (operating conditions: Tsup = 90 °C.drain to leave the outer film.i Ri Fi S R Ri Fi dAi (2) Ai hi 2 The residual of the drag torque balance is reduced with an iterative Newton- Raphson algorithm by adjusting the floating ring rotational speed. film thickness (ho.o which is taken as a boundary condition in the outer films’ Reynolds equation. Fi). The feed pressure in the inner film pl. ηi). With full floating rings. A conservative finite difference scheme (FVM) is then used for solution of Reynolds and energy differential equations [14]. 3 EXPERIMENT For model validation. For ring speed measurement. the lubricant is supplied to the inner film through six equally spaced radial feed holes in the full floating ring. The set-up of the hot-gas turbocharger test rig and the instrumented turbocharger is shown in Figure 3. 90. Table 1 . This is taken into account by the thermal model by calculating the mean temperature of the lubricant leaving the outer film through the drainage area and feeding the inner film with the same temperature.Rotor and bearing system data Rotor length l = 100 mm Rotor mass m = 100 g Lubricant dynamic viscosity at 40 °C η = 71. The fluid entering the holes on the ring’s outer circumference is preheated due to heat flows into the outer film (from TC casing surface and through the floating ring) and shear-driven power loss. To obtain realistic experimental data as model input and for validation purpose. A passenger car turbocharger with full-floating ring bearing is used (Fig.drain. thermal boundary conditions have a substantial impact on bearing performance. the equilibrium between the outer static load and the film reaction forces is calculated. Table 1 summarizes the relevant rotor and oil supply information. experimental investigations under varied lubricant supply conditions are conducted in a turbocharger test facility of the Institute of Turbomachinery and Fluid Dynamic of Leibniz University. 110°C As described in the preceding section. In the outer loop of the computational program the floating ring speed is altered which requires the beforehand listed steps to be repeated until the thermal state and the drag torques on the floating ring surfaces are balanced. The lubricant is fed to the outer film through a half-moon shaped groove spanning one third of the casing surface circumference.i pl . Additionally. The energy equations of the films and the heat conduction equations for the solid bearing sections are solved in a single linear system of equations. an eddy current turbocharger speed measurement system is used for both FRBs. 3). The hybrid scheme is utilized to stabilize the conjugate heat problem of the energy equation [18]. the turbocharger is instrumented with about forty temperature and thirty pressure probes.drain equals the total lubricant flow rate leaving the inner film region Qi. The sensors are instrumented in the bearing casing tangentially to the end of the individual oil supply groove detecting the passage of the radial feed holes in the floating rings. From the outer film. 426 .o pc (4) is adjusted by an iterative algorithm until the amount of oil leaving the outer film region through the specific drainage area Qo. The numerical procedure starts with estimated values for temperature. the compressor ring is instrumented with a high-frequency “Kulite” pressure sensor detecting the time- resolved pressure in the oil supply groove. In the first step.pl . The thermal state of the FRB is solved with subsequent adjustment of the operating clearances. pressure and viscosity distribution and an assumed floating ring speed.98 mPas Lubricant supply pressure psup = 3 bar(abs) Lubricant supply temperature Tsup = 70. This qualitative behavior was also shown by several authors for measurements with a cold gas driven turbine and an oil supply temperature of 37. the outer temperature boundary conditions are nearly constant at low rotational rotor speeds mainly affected by the supply conditions.Ring speed measurement (left).000 rpm in steps of 10. schematic of the floating ring temperature instrumentation (right) The experimental investigations are conducted with the lubricant type SAE 5W-40. In the present case. instrumented turbocharger (right) The bearing brackets are instrumented with four thermocouples at the turbine side and two at the compressor side at three axial positions. The test matrix includes three lubricant inlet temperatures and the shaft speeds tested ranging from 40. The turbine inlet temperature is constant at 600°C for all test runs. The shaft temperature affecting the inner oil film temperature is a function of viscous dissipation leading to a stronger decrease of ring speed at higher rotor speeds.7°C [19]. Sensor Figure 4 .000 rpm to 200. 5 (a)). Measurements with hot gas driving the turbine show a difference in speed for the compressor and the turbine radial bearings at TC part-load condition due to heat flows. Figure 3 . The maximum speed ratios are measured at the highest lubricant feed temperature decreasing with higher shaft speed. three temperature probes have been placed in the upper section of the bearing housing to detect the global temperature gradient from the turbine to the compressor side.000 rpm. The compressor operates at the optimum efficiency of each speed line shown as red line (Fig.Set-up of hot-gas turbocharger test rig (left). Figure 5 (b) displays the measured operating points and floating-ring speed ratios Ω (rotational ring speed divided by shaft speed) of the compressor and the turbine ring for the three different oil supply temperatures at constant oil inlet pressure of 3 bar(abs). The 427 . In addition. Figure 5 (c) shows the compressor outlet temperature and the bearing housing temperatures at three different positions. Figure 4 gives a schematic sketch of the floating ring instrumentation. Figure 6 (c) depicts the orbit measurement for different rotor speeds. For a rotor speed of about 1800 Hz. resulting in a change of the effective viscosity and ring speed ratio. The corresponding dynamic pressure in the compressor ring oil feed sickle is shown in Figure 6 (b). Above rotor speeds of 2800 Hz. At 1000 Hz. In addition.000 krpm related to the subsynchronous 1-3. the temperatures and thus the rings speed ratios converge for both rings. averaged bearing bracket temperatures (d) At part-load and depending on the oil supply temperature.turbine side shows a nearly linear characteristic whereas the compressor side follows the compressor outlet temperature. bearing housing temperatures (c). The corresponding bracket temperatures are shown in Figure 5 (d). the shaft temperature rises to higher compressor pressure ratios and dissipation in the inner fluid film which leads to a further increase of the shaft temperature. the 428 . The test data shows three significant subsynchronous motions up to shaft speeds of 180. ring speed ratios (b). the inner fluid film becomes stable and the oil whirl/whip in the stable outer fluid film becoming unstable excites the gyroscopic conical forward mode (Subsynchronous 3).Measurement results of operating points (a). the inner casing surface acts as a heat source or sink and the heat flow from the casing affects the mean temperature of the outer film. In the rotor speed range from of 500 to 1000 Hz. the rotor vibrates in the gyroscopic conical forward mode (subsynchronous 1). Figure 6 (a) shows a waterfall plot and the orbit for the shaft displacement for rotor speeds from 40 to 200 krpm. (a) (b) (c) (d) Figure 5 . the system jumps into the gyroscopic translational forward mode (subsynchronous 2). For high rotational speed and high dissipation in the bearing. For the high amplitudes in the shaft speed range from 100 to 190 krpm.006 nFR. subsynchronous 3 disappears and the rotor vibrates around a stable equilibrium position [20].04 2*sync noil feed holes sync 0.01 nrotor 0. The subsynchronous 3 corresponding to ≈50% of the ring speed denoting instability of the outer FRB oil film. 21.T 0.08 nFR.T amplitude in bar 0.5* sync nFR 1*sync nFR Sub 2 1*sync nrotor 0.004 0.02 4*sync noil feed holes 0 200 rot o rs 150 10000 pe 8000 ed 6000 in 100 kr p in Hz 4000 m 2000 ency frequ 50 0 Figure 6 .C 0.Waterfall plot of shaft displacement (a). waterfall plot of dynamic pressure (b).C 1*sync noil feed holes 0. orbit measurements for different rotor speeds (c) These results matches quite well the rotordynamic behavior presented in previously published literature[20. subsynchronous 3 increases 429 .008 amplitude in mm nFR. 22].06 nFR.002 Sub 3 0 200 Sub 1 ro t o rs 150 1500 pe ed 1000 in 100 kr p 500 in Hz m ency 50 0 frequ (b) 0. (a) (c) 0.1 nrotor 0. In both cases. Due to the fact that a shaft temperature measurement is not practical. In both cases. the mean temperatures of the bearing components surfaces and of the fluid films are close to the system supply temperature at low shaft speeds. only the results of the turbine-side FRB are shown.Predicted mean temperatures of lubricant films and solid surfaces In both plots of Figure 7. the average film and ring surface temperatures stays at levels slightly above supply temperature even at high shaft speeds. The oil supply temperature is 90°C and the lubricant is fed to the outer film with an inlet pressure of 3 bar(abs). 430 . For simplicity. the casing temperatures measured on the hot gas test rig are used as outer thermal boundary conditions for the casing segment to gain a better comparability. the inner fluid film exhibits a high temperature gradient which exists in radial direction at high shaft speeds. The shaft surface is actively cooled by the inner film lubricant flow resulting in an inhomogeneous temperature distribution in the shaft segment. For the externally heated shaft (right) this characteristic is found less prominent. The characteristic of the inner ring surface temperature appears linked to the inner film temperature. As for the adiabatic shaft. the shaft end temperatures are assumed to be 200°C. The predicted results for an adiabatic shaft segment (no heat flow in or out of the shaft) are similar to those results expected for a turbocharger driven by cold gas. Figure 7 . 4 COMPUTATIONAL PREDICTIONS A heat flow analysis for two different thermal boundary conditions of the shaft is presented below.leading to higher eccentricities of the ring and thus to a decreasing ring speed [23. the temperature gradient between the inner and outer ring surfaces increases significantly with higher shaft speeds. A variation of the predicted mean shaft temperature surface across the operating range is found even for the model with constant temperature at the shaft axial cut faces. For the actual predictions this temperature is used as a boundary condition on the axial cut faces of the shaft segment considered by the model. A deeper rotordynamic analysis based on the measurements presented in this paper is given by Vetter [23]. The results calculated with a uniform temperature prescribed at the cut faces of the shaft give a fair approximation for the operating characteristic parameters of a turbocharger under realistic operating conditions driven with hot exhaust gas. 20]. The most relevant predicted temperatures are shown in Figure 7 which differ significantly for the two cases. For the adiabatic shaft model (left). The same characteristic of decreasing compressor ring speed is shown in [24]. Prescribing a uniform temperature at the shaft outer faces. a small heat flow from the casing into the outer film is also found. As the measured casing temperature lies above the lubricant inlet temperature throughout the whole operating range and slightly increases towards the maximum shaft speed.e. normalized with the maximum heat flow from the shaft. In case of an adiabatic shaft (top row).Figure 8 depicts the dimensionless heat flows within the FRB plotted over the operating range. the shear-driven drag power loss. there is only a limited amount of heat that can be carried away convectively. This leads to an increasing temperature gradient over the ring in the radial direction. Due to the excellent thermal conductivity of the floating ring material (brass) and the high radial temperature gradient. Even at low shaft speeds. The inner film mean temperature starts at the outer film’s supply temperature at low shaft speeds but heats up with increasing speed. a second heat source is introduced. a significantly different heat flow characteristic is found. i. by contrast stays near the supply temperature even at high shaft and ring speeds due to the large amount of oil flow. Here the heat flow mixes with the heat dissipated due to surface shear and the pressure driven flow generated by the ring rotation. As the lubricant flow rate through the inner film is very small compared to the outer film flow rate. the larger portion of heat is conducted through the floating ring into the outer film.Top row: Dimensionless heat flow characteristics with adiabatic shaft. the inner film mean temperature reaches high levels above the 431 . Figure 8 . bottom row: with shaft cut faces at a prescribed temperature of 200°C For the non-adiabatic approach (bottom row) especially for the inner film. the only heat source in the inner film is dissipation. The outer film. In the outer film. the total amount of heat is carried away convectively with the oil flow. the heat flows similar to the adiabatic shaft case. Figure 9 . the outer film mean temperature rises and the pressure drop in the ring feed holes grows resulting in a reduced flow that enters the inner film trough the feed holes at higher temperature. the ring speed ratio shows only a slight decline over the operating range. The predicted heat flows. entering the inner film through the feed holes is – although it has been preheated in the outer film – considerably colder than the inner film fluid and thus carries heat out of the inner film. especially for the inner film region. the mean viscosity starts at a high level close to the outer film viscosity but drops rapidly with increasing shaft speed. the characteristics of the predicted ring speed ratio for both thermal boundary conditions differ significantly. the radial temperature gradient through the ring increases and leads to an enhanced conductive heat transport. For the adiabatic case. except for the heat flow entering through the ring outer surface which is at a high level even for low shaft speeds already. With the shaft acting as a heat source. the inner film viscosity already starts at a low level for small shaft speeds and reduces further with increasing shaft speeds but at a significantly lower rate compared to the outer film’s viscosity characteristic.Dimensionless mean effective lubricant viscosities (left) and ring speed ratio for varied thermal boundary conditions (right) The predictions show a significant influence of the shaft thermal boundary conditions on relevant bearing parameters like temperature and lubricant viscosity distribution. the adiabatic shaft model leads to a decreasing ring speed ratio while for the heated shaft. As the ring speed increases. differ considerably. The shaft temperature has only limited influence on the outer film properties due to the high local lubricant flow. the convective heat flow (lubricant carries heat away) in the inner film region is dominant over the conductive heat flow through the floating ring. Figure 9 depicts the influence of the shaft thermal boundary conditions on the mean effective lubricant viscosity in the inner and outer film regions shown normalized with the lubricant viscosity at the oil inlet temperature (90°C). mechanical power dissipation. The fluid. In the outer film region. a similar characteristic is found in both cases.outer supply temperature. Due to the mean viscosity gradient between the inner and outer films developing with increasing shaft speed. At low shaft speeds. A different characteristic of the mean lubricant viscosity is evident for the inner film. For the outer film. and ring speed ratio. The resulting lower mean effective viscosity leads to a decreased drag torque on the floating ring inner surface which consequently lowers the ring speed ratio compared to the adiabatic shaft case. 432 . As the ring speed is primarily determined by the ratio of the mean viscosity between the inner and outer film. While the heat is carried away by lubricant convectively. These results may vary for different operating conditions like modified oil feed pressures and temperatures or exhaust gas entering the turbine with higher temperature or active cooling of the TC casing. since the oil flow in the outer film is significantly larger than in the inner film. This leads to the conclusion that shaft temperatures during the operation are higher than 200°C such that a reduced viscosity and thus drag torque is present in the inner film. predicted ring speed ratios are compared with measurement results from the hot gas test rig (Fig. Predicted ring speed ratios for an oil supply temperature of 110°C are higher than measured ring speed ratios throughout the whole operation range. for the studied TC under the stated operating conditions. the prescribed shaft temperature leads to an overestimation of the viscosity reduction in the inner film because the heat flow from the shaft results in too low a prediction of the ring speed ratio. 5 COMPARISON WITH HOT GAS TEST DATA For model verification. with the adiabatic analysis misjudging the real ring-speed ratios significantly. the heat flow from the casing increases the mean temperature of the outer film resulting in a reduced effective viscosity and a (slight) decrease in ring speed ratio.Predicted ring speed ratios compared to test data for varied lubricant supply temperature (left). However. The expected actual shaft 433 . three discrete shaft speeds and shaft temperatures ranging from 100°C to 250°C in steps of 25°C. Figure 10 . At shaft speeds below 60 krpm. varied shaft thermal boundary conditions compared to test data for supply temperature at 70°C (right) To get an indication of real shaft temperatures during test rig operation. left). sensitivity analyses were performed for a lubricant supply temperature of 70°. even though the trend is matched well. This indicates that the shaft temperature is lower than 200°C in reality. The measured TC casing temperatures are used as thermal boundary conditions for the casing model. For oil supply temperature set to 70°C and 90°C the predicted ring speed ratios show an excellent correlation with test data at medium to high shaft speeds.The casing thermal boundary conditions influence the heat flow and the bearing operating conditions as well. the extra heat from the casing surface is carried away efficiently and thus only leads to minor variation of the bearing parameters. The results are shown in Figure 10. the results predicted for an adiabatic shaft segment are shown as well (gray). If the inner casing surface acts as a heat source (as it does for the TC studied). For comparison. The shaft temperature is again set to 200°C for all calculations. 10. For brevity. only results from the turbine-side FRB are shown. C. A sensitivity analysis of the shaft temperature revealed the serious impact of the shaft thermal boundary condition on the bearing operating parameters.. M. The present work shows that the knowledge of the thermal boundary conditions is mandatory for a precise prediction of the bearing performance parameters.. V. D. F..temperature can be obtained by searching for the intersections of the measured ring speed ratio curve with the ring speed ratios calculated for different shaft temperatures at the particular shaft speed. a complex thermo-hydrodynamic computational model for the prediction of relevant FRB operating parameters was presented. “Whirl and Friction Characteristics of High Speed Floating Ring and Ball Bearing Turbochargers”. For making realistic assumptions concerning the thermal boundary conditions for the FRB and for model validation purposes. The computational model was used to study the thermal boundary condition influences on bearing operating parameters and heat flow characteristics. (2012). 041102 [2] Shi. J. J. J. In contrast. the bearing simulation model should be integrated into a TC performance prediction model including non-adiabatic approaches to determine thermal boundary conditions for the numerical bearing simulation. “On the Effect of Thermal Energy Transport to the Performance of (Semi) Floating Ring Bearing Systems for Automotive Turbochargers”. SAE Technical Paper 2011-01-0375 [3] San Andrés. “An Analysis for Floating Bearings in a Turbocharger”. Bielenberg. Comparing predicted ring-speed ratios assuming an adiabatic shaft with test data from a hot gas test bench showed a significant discrepancy. For further improvement of the prediction quality.. for his crucial contribution to instrumentation and experiments. D. Deng. 6 CONCLUSIONS In the present paper. A. Barbarie. Archer. an extensive experimental study on a hot-gas turbocharger test bench was conducted. (FVV) whose funding is gratefully acknowledged. of Tribology. 135(10).. particularly in TC part load conditions.. Sadeghi. the thermal influence of the thrust bearing as well as water cooled bearing housing should be considered. Mr. predictions produced with a fixed shaft temperature showed a significant improvement in agreement with the test data over a wide operating range. Specifically. of Engineering for Gas Turbines and Power.. Donaldson. 134(10). ASME J. C. V. F. K. Bhattacharya.. (2013). (2011). The authors would also like to thank Borg Warner Turbosystems GmbH for substantial material support during the study. 102507 434 .. L. ACKNOWLEDGEMENTS The work was supported by the Forschungsvereinigung Verbrennungs- kraftmaschinen e... a significantly lower shaft temperature of 150°C than assumed is found while at high shaft speeds the assumed 200°C seem to fit the experimental results well. In addition the authors would like to thank the technician. Neglecting non-adiabatic influences leads to an inaccurate prediction of TC performance in rotor dynamic and combustion engine process simulations. 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Seume.: “Dynamics and Stability of Automotive Turbochargers”. K. and LaRue. 1035-1046 [20] Schweizer. “Thermal Effects on the Performance of Floating Ring Bearings for Turbochargers. ASME Paper No. 2004. T. 129(4).. 2003.. pp... Dusseldorf. LaRue. San Antonio. H.. Chinta..” Proc. Kukla. C. VDI series 1 no. Archive of Applied Mechanics... P. pp. Heuer. .: “Rotor Dynamic Analysis of a Passenger Car Turbocharger using Run-Up Simulation and Bifurcation Theory”.. (2013). Seemann. E. IMechE. London.. Schwarze. “Predictions for run-up procedures of automotive turbochargers with floating-ring bearings including thermal effects and different bearing setups”. of the 11th Int. IMechE. “Entwicklung Eines Abgasturboladers für die Schweren Nfz-Motoren von Daimler”. 321. J. pp.. Boyaci.. Müller.10. Gorbach.jsv. Leweux.. Proppe. W. D. 9th International Conference on Turbochargers and Turbocharging. B. Sievert. GB (to be published) [24] Chebli..013 [22] Tomm. Doi: 10..2008. Issue 2. Journal of Sound and Vibration. on Turbochargers and Turbocharging..: “Nonlinear Oscillations of Automotive Turbocharger Turbines”. Vol.. M. M. pp. May 2010 [23] Vetter. Esmaeil.[21] Schweizer.. Busch. T.1016/j. Proc.. London. pp 124-129 436 . 955- 975.. Hagemann. 335-347. Schweizer. H. L. Conf. B. MTZ - Motortechnische Zeitschrift. G. M. 2009. (2014). Volume 74. A. C. U. Stability analysis of pressurized Gas Foil Bearings for high speed applications R Hoffmann. A multi physics model is used to calculate dynamic properties (stiffness and damping). 2014 437 . Due to friction contacts inside the corrugated structure a structural damping is induced and can be estimated experimentally (3)(4). In addition Kim et al. recent developments in GFB applications have heightened the need of an appropriate thermal management for high temperature environments due to limited temperature durability (7)(8). Berlin Institute of Technology. In addition the effect of a pre-swirl side feed pressurization is considered. using multiple bump layers results in higher load capacities and can reduce high sub synchronous whirl amplitudes (2)(5)(6). Therefore an optimal film thickness is achieved and higher loadings compared to rigid gas bearings are possible (2). Germany ABSTRACT One of the most efficient cooling methods for Gas foil bearings is a side feed pressurization. R Liebich Chair of Engineering Design and Product Reliability. Low drag friction. An elastic structure comprises one or more thin top foils supported by corrugated bumps (see Figure 1a) which are the main difference to common gas bearings with a rigid bearing housing. Besides cooling effects. One of the most efficient cooling method is a side feed pressurization of GFBs (7)(9)(10). applying coatings. Hence. This current paper examines the effects of side feed pressurization on the dynamic behaviour. high speed operation and the omission of an oil system are some advantages of compliant foil bearings (1). A tuning of the compliant structure by staggering of bumps. _______________________________________ © The author(s) and/or their employer(s). which have a major impact on the rotor dynamic performance. Nevertheless. This pressure is induced by a generated slip stream between the turning bearing journal and the bearing foil (see Figure 1b). Heat is transported by convection and conduction effects. The sub synchronous whirl vibrations have a major impact on the dynamic behaviour. experimental investigations have shown a reduction of sub synchronous vibrations. (10) have shown that the effect of this method reduces the sub synchronous amplitudes for low loaded bearing conditions (W≈ 5 N). In general. GFBs are based on the hydrodynamic pressure. 1 INTRODUCTION Over the past three decades Gas foil bearings (GFBs) have been successfully introduced in the field of high speed turbo machineries. T Pronobis. To reduce sub synchronous vibration several methods and devices have been introduced (11)(12)(13)(14). special coatings and cooling techniques have been established to prevent thermal instabilities and thermal seizure (8)(9). A forced cooling flow is streaming underneath the corrugated bearing structure and the clearance between bearing journal and top foil (see Figure 2a). The compliant term ℎ ( . ) and the reacting force vector fB is zero. The first part includes the nominal bearing clearance and the journal centre displacement. where ( ) is the pressure underneath the foil and along the axial direction z.However. which is discretized by a hybrid finite difference scheme. ) is a) b) Figure 1. (1)). It acts under the attitude angle = ( . Finally. The equilibrium condition is reached if the sum of the loading vector = ( . The film thickness is composed of two parts: A rigid term ℎ ( ) and a compliant term ℎ ( . as shown in (11). where and are related to the axial and circumferential coordinate. and ) is shown in Figure 1b). It is limited to small orbits around a static equilibrium position. Schematic GFB structure a) and GFB with a dynamic pressure field b) induced and results in a dynamic pressure field ( . However. (2) describes a perfectly aligned journal. ). A perturbation method is applied to estimate linearized bearing parameters. expansion effects due to temperature gradients and centrifugal forces are neglected. a 2D FE-model of the corrugated structure is coupled with the non-linear compressible Reynolds equation. 2 THEORETICAL MODEL The elastic structure of a simple GFB in a coordinate system ( . Due to the presence of a turning journal with an angular speed and centre displacements εx and εy a forced slip stream with a film thickness of ℎ( . ( . there has not been a discussion about the impact of side pressurization in detail and especially the impact of a pre-swirl pressurized GFB on rotor dynamic stability has not been reported. they are inputs of a linear stability analyses. An integration of the pressure field along the axial and circumferential direction yields the reacting force vector fB (Eq. ) describes the deformation of the corrugated structure due to the pressure field and is calculated by a structural model as shown below. to estimate the onset speed of whirling motions it is an appropriate method. A pre-swirl has a direct influence on the velocity field inside the film thickness of a GFB. First. )− ( ) cos =− d d (1) sin Eq. ). This pressure generates a reacting force fB. linearized bearing parameters will be analysed in detail. ). 438 . Eq. Note the rotation speed is a function of due to side feed pressurization. ) is calculated by solving the Reynolds equation for a compressible. ( )= 1− + (5) Where = .1 Side feed pressurization a) b) Figure 2. Side feed pressurized GFB a) and swirl generator b) Figure 2a) shows a schematic side feed pressurized GFB. (4). The pressure of the leading and trailing edge of top foil at = 0° is given by ( = 0. ) = ( = 2 . Eq. Due to a side feed pressurization the circumferential speed becomes a function of and can be described by Eq. A pressure drop underneath the top foil along the bearing length is generated due to the pressure difference between the feed pressure and ambient pressure . ) = . (5). = (6) (Temperature and ideal gas constant ) 439 . ( ) + = ( ℎ) + 12 ( ℎ) (3) 2. 0) = and ( . It has been used by Black et al. (3). However. while neglecting swirling ( = 0). isothermal and isoviscous fluid. ) = + cos θ + sin θ + ℎ ( . ℎ( . It links the pressure field with the film thickness under the presence of journal rotation speed ( ). (10) have applied this equation to estimate side feed pressurization of GFBs. ( )= 1− + (4) Following boundary conditions are applied for solving the Reynolds equation: ( . ) = ( ). θ) (2) ( ) The pressure field ( . where R is the bearing radius and μ the absolute viscosity. (15) to describe pre-swirled gas seals. y. Eq. and = = . based on Iordanoffs (16) bump model. The overall structure can be explained by a finite element model. ) ). ) (12) 440 . The complex structure is reduced towards a 2D Timoshenko plate for the top foil and a linked spring model. (8). the swirl number will depend on the swirl angle ± and the pressure ratio ∏ = . These guide vanes are placed at one side of the gas foil bearing. Schematic structure a) and model including sagging b) In Figure 3a) a schematic compliant structure of a GFB is shown. Figure 3b). It is the ratio between the nominal circumferential velocity = and the circumferential part of the absolute velocity of a pre-swirl guide vane.2 = (7) Due to neglecting of boundary layer effects in the swirl generator an isentropic flow can be assumed and yields to Eq.4.) a perturbation method is applied. as given by Eq.2 Structural model a) b) Figure 3. the pressure field and the film thickness are affected by the perturbation. as shown in Figure 2b). 2. In addition. . To avoid sagging effects the Young-modulus of the plate is increased by a factor of 4. (14). are given.. This method was introduced by Lund (17) first. . = +∆ (10) = +∆ (11) ℎ = ℎ + ∆ (ℎ + ℎ . where i=x. ≪ ) acting around the equilibrium state ( = ( . If = = . ℎ ⋂ = (9) Where the pressure force vector is based on the pressure difference of ( . ref. .3 Dynamic performance To estimate the dynamic bearing parameters for a given load w and steady speed condition (Ω =const. where =1.Considering a pre-swirl. (9). ) − ( ). The compliant part ℎ is a subset of the displacement vector . −1 ( ±) 2−1 1 1− ∏ = . (11) and (12). Assuming a harmonic perturbation with the perturbation frequency and small amplitudes (∆ . the swirl number is applied.2 ≡ (8) 2. = . = 1 bar) and a nominal clearance c of 31. ) and ( . and the critical modal stiffness .2 is considered.Where ℎ = cos and ℎ = sin . (13). analogously to Eq. (14). cos cos + =− d d (14) sin sin Note the described model has been validated by experimental data of Ruscitto et al. a higher node density with an axial element length of = /120 is introduced for 5 axial nodes at each 441 . Positive real parts indicate an unstable behaviour of a linear system. (17)(18). due to the dependence of the bearing parameter. (3)) while neglecting terms of higher order leads to a zero and a first order Reynolds equation. while damping of the top foil is neglected. ref. is the perturbed compliant film thickness. . = 2 (16) . Finally. 3 NUMERICAL RESULTS Table 1 lists the geometrical data of the investigated GFB. to introduce viscoelastic damping effects of the bump structure. see Eq. Using the perturbed pressure field to calculate the linearized stiffness and damping matrix. It is solved by using a harmonic state of solution and assuming non trivial solutions. For a detailed description see ref. these parameters can be used for introducing the critical mass parameter to evaluate stability (11). 0 0 + + = (15) 0 0 Introducing the impedance ℎ = + and considering a free oscillation yields to an eigenvalue problem. (9). Simulations are based on ambient conditions ( = 293.15 K. The term ℎ . It is based on the perturbed displacements. ). The mesh grid includes 130 circumferential and 40 axial nodes. = 3. Different bump stiffness factors are applied. calculated with the model of Iordanoff (16) ( . The eigenvalue is defined by = ± . For evaluating the dynamic performance of a GFB system a simple linear oscillator model is introduced. which has been widely used to analyse GFBs. where is the journal mass.87 ∙ 10 N/ mm ) and a structural loss factor of γ=0. . ref. (17)(18). The global stiffness matrix is changed to a complex stiffness matrix. It enables the calculation of the critical excitation frequency . (19) for a non pressurized case. = 8. To determine the neutral stability for a steady speed condition ( = . The structural viscoelastic damping is related to the dissipated energy caused by friction contacts inside the compliant structure and can be described by the structural loss factor γ. A hybrid finite difference scheme is applied to calculate the pressure field ( . Further.) the excitation frequency ratio is varied.8 μ is assumed.76 ∙ 10 N/ mm and . = + (1 + ) (13) Substituting the perturbed approaches (Eq. Note for neutral stability the modal damping is zero. For = 0 the neutral stability is gained. (10)-(12)) into the Reynolds equation (Eq. 1016 mm bump pitch 4. Table 1: Geometrical data of a GFB (17) Parameter Variable Value Unit bearing radius 19.8 μm foil thickness 0. Note.0E+03 1.0E+04 15 N unstable region 1. They are related to the journal mass and correspond to the loading ⁄ ).edge along circumferential direction to achieve higher numerical stability due to huge pressure gradients in this area. in the first analysis no feed pressure is supplied. Whereas the stiffness difference of the cross coupled stiffness ( ) is slightly reduced for a sub synchronous excitation ( 1).07 ∙ 10 N/mm² Poisson ratio 0.27 - 1. As shown in the diagram increased loads lead to a higher stability range while higher speed numbers reduce the range of stability.1016 mm Young modulus 2. Note 442 .0E+06 1. 15 and 30 N.0E+01 5N 1. while the excitation frequency ratio ⁄ was varied between 0. An increase of mass parameter or rotor speed beyond this threshold will cause an unstable behaviour. Three different side feed pressures were observed (1. Critical mass parameter vs.1 mm bump thickness 0.05 mm bearing length 38. The area underneath each curve is called stable region. The critical mass parameters with values of 10 are stable and hold for an adequate logarithmical display.778 mm clearance c 31.0E+05 critical mass parameter 1. 1.0E+00 6000 12000 18000 24000 30000 36000 42000 rotor speed rpm Figure 4. Furthermore. three horizontal dashed lines are plotted. rotor speed without side feed pressurization For a set of loadings (5.0 bar).100 rpm for the loading cases 5.800. As reported in (14). 15 and 30 N) the critical mass parameter versus the rotor speed is shown in Figure 4.0E+02 1.500 and 22. higher loading along the vertical direction raises direct stiffness and damping values .572 mm bump number 26 - half bump length 1. The onset speed is estimated at 14. 16.5 and 3.0.1-10.0E+07 30 N p=1bar 1. has a major impact on the stability: Small values of results in less coupling effects and can raise the range of stability (20).). influence of a side feed pressure (without swirl = ) Figure 6 shows bearing parameters of three different feed pressure levels 1. High values can decrease the range of stability. whereas the 1 bar values are based on a GFB without side feed pressurization (ambient condition).0 bar 1.9 N loading. However. In addition. as shown in Figure 5. different levels of side pressure without a pre-swirl are listed. Only for 30 N and speeds of 6. Furthermore. All values are displayed versus the excitation frequency ratio ( = ⁄ ).5 bar p=3. It should be mentioned. excitation frequency ratio and feed pressure yields higher direct stiffness. rotor speed. Those effects of pressurization on the direct parameters are more notable for low rotor speeds. while damping decreases and stiffness is conversely affected. 30 N = 1. This work was extended to 15 N and 30 N 443 .000 rpm the GFB shows lower stability ranges for side feed pressurization. an increase of rotor speed or excitation frequency ratio significantly drops the direct damping values.0E+02 15 N 1. As mentioned before. Critical mass parameter vs. The damping is higher due to feed pressures compared to the case of p=1 bar. In addition. that similar results for and were observed. The difference rises with increasing rotor speed and decreases towards zero for higher excitation frequency ratios ( 1.000-20. due to the higher damping and stiffness values for a side feed pressurization.000 rpm for a medium loaded bearing (30 N). for a sub synchronous excitation range the difference of the cross coupled stiffness is bigger for a side feed pressurized condition. similar results have been reported by Kim et al. the work of Kim et al.5 and 3 bar. the overall behaviour results in lower stability in the speed range between 6. Increasing side feed pressure raises the range of stability slightly. However. higher excitation frequency ratios and raised rotor speeds affect damping and stiffness. This effect is not notable for loadings of 5 and 15 N. Hence. 1.0 bar 5N p=1. As shown in the figure.0E+00 10000 15000 20000 25000 30000 35000 40000 rotor speed rpm Figure 5. (10) was limited to 4.0E+01 p=1. for 5 and 15 N. The GFB is loaded by a vertical load of 30 N. compared to the case p=1 bar. The same effects as mentioned above due to loading and rotor speed are plotted. the difference of the cross coupled stiffness has a major impact on stability.000-23. In Figure 5 the critical mass parameter versus the rotor speed is shown for the same loading set. increasing rotor speed.0E+03 critical mass parameter 1. (10). 6.000 rpm p=3 bar.E+06 p=1 bar. 10.E+06 6. 10.E+00 higher rotor speed -1.E+03 direct damping cxx in N/m 4.1 1 10 excitation frequency ratio η Figure 6.E+03 0.E+06 1.E+06 30 30NN 5.000 rpm p=2 bar. The numerical results of the analysed bearing design indicate no benefit of a side feed pressure on the stability for 30 N bearing loading.000 rpm cross-coupeld stiffness kxy-kyx in N/m p=2 bar.000 rpm p=3 bar. W=30 N) 444 .E+00 3. a GFB at ambient conditions (without swirl = .000 rpm 2. 70. 70.E+06 higher rotor speed 3.E+06 0.(medium loaded conditions). 70.E+06 direct stiffness kyy in N/m 4. 10. Bearing parameters at different levels of feed pressure vs.E+06 p=1 bar.E+06 2.E+03 higher rotor speed 2.E+06 0.000 rpm 1. 5 and 3 bar.5bar 15N.16 are used. This effect is more notable for higher pressures and loadings. p=3.8 30N.5 bar) the swirl increased the mass parameter up to 3.2 0 0. = 0).0bar 30N.2 increased 15N 5N.6 15N. However.4 30N 5N. A swirl angle ± = 20° and two different pressure ratios Π = 1. at high speeds the pre-swirl of positive and negative swirling angles decreases the range of stability. and indicates the increased cross coupling effects. The simulations are based on the same side feed pressures as mentioned above. ) are bigger than the nominal clearance. 0 5N rotor speed 0.6 horizontal excentricity ey Figure 7.33 and 1.0bar 15N.4 0.0bar 0.57 %.3 0. p=1.5bar 30N.5bar loading 5N. Figure 8 shows the relative deviation of the critical mass parameter related to the swirl free state versus rotor speed. = . p=1. p=1.0bar 0. which is possible due to the compliant structure. p=3. Finally. a pre-swirl is applied for the second part of this investigation. For increasing load (acting in vertical direction) the displacements are raised and dimensionless eccentricities ( = ⁄ . An increase of rotor speed leads the centre of mass towards the centre position ( = 0. while higher side feed pressure yields to higher eccentricities along the y axis. Influence of side feed pressure on the static Trajectory (without swirl = 0) Figure 7 shows the static trajectories of the static equilibrium state for different loading and rotor speed conditions versus the different side pressure levels. At low speed conditions and 15 N (ps=1. p=1bar vertical excentricity ex 0.2 0. due to feed pressure. p=1. 445 .5 0. The pressure ratios are related to the pressures ps=1.0bar 1 nominal clearance 1. p=1.1 0. p=3. Further. Increasing pressure results in higher direct stiffness and damping.0bar.. pp.57 15N. alpha+ 3. The current investigation was limited on linear stability. (3) Ku. Journal of Vibration and Acoustics. p=3.000 rpm. alpha- 1 30N.0bar.5bar. 287–294. p=1. H.5bar alpha- 4 15N. Relative deviation of the critical mass parameter with swirling ࣘ ് related to the swirl free case ࣘ ൌ 4 CONCLUSIONS This paper has presented a model to investigate the dynamic performance of a gas foil bearing affected by a side feed pressurization. p=3. 116(2). (2000). and Heshmat. pp. H. a test rig is under construction at the Berlin Institute of Technology to estimate linearized bearing parameters and dynamic performance of different side feed pressure levels to prove theoretical findings.. p=1. However. (1994). alpha+ 30N.0bar. Advancements in the performance of aerodynamic foil journal bearings: High speed and load capability. p=1. p=3.5 bar side pressure and low rotor speeds <10. due to higher amounts of damping the overall stability has been improved for loadings of 5 and 15 N.0bar. M. alpha- 3 5N. alpha+ 5N. (1994). Effect of static load on dynamic structural properties in a flexible supported foil journal bearing.000 rpm the stability has been improved.5bar alpha- 0 -1 -2 -3 decreased stability range -4 -5 4000 24000 44000 64000 rotor speed in rpm Figure 8. 15N. Journal of Tribology. In addition. p=3.5bar alpha- 2 raised stability range 5N. p=1. p=1. future research will be based on time transient simulations to confirm the impact of a side feed pressurization on the rotor dynamic performance including nonlinear friction contacts. while cross coupling effects have been increased for higher loadings and side feed pressures. (2) Heshmat. deviation ∆߰ܿ ݐ݅ݎin % 5N. C. alpha+ 5 15N. alpha+ rel. Only for loadings of 15 N and 1. 257–262. 446 .5bar. pp. 116.5bar. 795–801. p=1. C.-P.. 5 REFERENCE LIST (1) DellaCorte. the results have shown that pre-swirling has slightly decreased the range of stability.. Load capacity estimation of foil air journal bearings for oil-free turbomachinery applications. and Valco. but loadings of 30 N have shown a stable behaviour only for high speeds >20. Inlet flow swirl in short turbulent annular seal dynamics. P. T. 135. (2011). Lee. Tribology International. H. (6) Heshmat. D. E. Valco. Journal of Lubrication Technology. Journal of Tribology. Mc Cormick. H. Joint Propulsion Conference. DellaCorte C.. (1991). W. and Choi..50:319–27. (11) Schiffmann. (1981). pp. pp. (14) Hoffmann... Technol. C. 121. Calculation of stiffness and damping properties of gas bearings. 447 . Foil bearing design guidelines for improved stability.... and Liebich. 131(1). W. Journal of Tribology. Proceedings of ASME Turbo Expo 2014. pp.(2009). S. S. (1982). L. H.. (1999). 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F. Nordmann. (2007). and Pfützner. Development of foil journal bearings for high load capacity and high speed whirl stability. and San Andrés. Parametric studies on static and dynamic performance of air foil bearings with different top foil geometries and bump stiffness distributions. Thermohydrodynamic analysis of bump- type gas foil bearings using bump thermal contact and inlet flow mixing models. Radil. T. dry friction coefficient. Analysis of an aerodynamic compliant foil thrust bearing: Method for a rapid design. and Kim. Analysis of compliant foil bearings with spatially variable stiffness... Z...(4) Rubio. LaMCoS. The Modified Phan-Thien and Tanner model takes into account complex rheological characteristics of the lubricant. transient aspects and inertia effects due to the very high rotational speed. It shows acceptable correlation between prediction and test data of axial thrust load versus thrust bearing oil film thickness for two typical turbocharger frame sizes. ) : Velocities [ ⁄ ] : Density [ ⁄ ] ℎ : Film thickness [ ] : Pressure [ ] : Stress tensor [ ] : Averaged flow rate [ ⁄ ] : Inertia terms [ ⁄ ] : Dynamic viscosity [ . France ABSTRACT In this paper. NOMENCLATURE . : Cartesian coordinates [ ] : Time [ ] ( . lubricants need to have appropriate viscosity to aid cold and start/stop functionality and need to have sufficient viscosity to provide enough capacity at high loads and temperatures _______________________________________ © The author(s) and/or their employer(s). 2014 449 . T Lamquina. An extensive rheological testing campaign was performed to get data as input for the numerical model. . : Models parameters [−] : Shear rate [ ] : Critical shear rate [ ] : Thermal conductivity [ ⁄ ⁄ ] 1 INTRODUCTION With increasing focus on engine downsizing. ] : Heat capacity [ ⁄ ⁄ ] : Temperature [ ] : Oil relaxation time [ ] . One of them consists in analyzing the phenomenon of “turbo lag” that delays the boost response and penalizes the customer drivability. . high performance and low fuel consumption. COE Shaft & Bearings.The Modified Phan-Thien and Tanner model applied to turbochargers thrust bearing B Rémya. CNRS INSA-Lyon. The results obtained from this model are compared to experimental data provided by a thrust bearing rig.b. turbochargers technology has to take up several design challenges to reach automotive engines expectations. B Bou-Saïdb a Honeywell Turbo Technologies. Engine lubricants are submitted to a wide range of temperatures. . France b Université de Lyon. an unsteady thermohydrodynamic formulation of a turbocharger’s thrust bearing contact is presented. pressures and shear rates. Although a low viscosity saves energy. The understanding of the bearing system load capacity and friction in steady state becomes not sufficient in extreme operating conditions and a transient formulation of the problem must be solved accounting for complex rheological characteristics and very high rotational speed effects. The first time step of the analytical model is compared to test data collected on a thrust bearing rig in terms of axial load and power loss. A specific software has been developed to solve numerically these equations. Some numerical parameters are based on an extensive testing campaign on a Couette rheometer. the ratio of the thickness over the radius varies between 10-2 and 10-3. The ideal lubricant should have a flat viscosity- temperature relationship. In addition. Note that ( ). But for mineral oils. Using Leibniz integral rule. 2 ANALYSIS Throughout this paper. yield to account for inertia effects.where wear might otherwise occur. Thus. All external forces are negligible and there is no slip of the fluid at the walls. the thrust bearing contact of a turbocharger is studied. we use thin film theory assumptions. a direct integration of the momentum equations is impossible. one obtains: + + = −ℎ + − [1] 0 = [2] + + = −ℎ + − [3] Where = = = = = Note that is the total deviatoric stress tensor. the flow is supposed incompressible and laminar in the region of contact where Reynolds number varies from 10 to 1000 being in the lower range of the laminar to turbulent transition. In this paper. it is to be expressed as the sum of a Newtonian stress and a polymeric stress based on the MPTT model. all second order terms such as (ℎ/ ) are neglected. up to 300 000 rpm. One of the most widely used rheological models that can represent those two lubricant properties is the Modified Phan-Thien and Tanner (MPTT) model (2). 2. ( ) and ( ) respectively are the tangential. Therefore.1 Modified Reynolds equation As inertia effects are considered. High rotational speeds. For the considered thrust bearings. In our study. radial and vertical directions of the contact. viscosity-index (VI) improvers in small percentage consisting in long-chain polymers are added to the solvent (1). These additives bring a satisfactory viscosity-temperature relationship but convey a non-Newtonian behavior to the lubricant. the viscosity drops with the temperature rise. The continuity equation is also averaged through the thickness: ℎ + + = 0 [4] 450 . The MPTT model is combined with two viscosity laws accounting for the shear-thinning effects (Cross model (3)) and the viscosity dependency on temperature (Walther and McCoull model (4)). we follow the method developed by Tichy and Bou-Saïd (5) which consists in averaging these equations through the film thickness. Therefore. − + = [8] 2 2 where tr = exp represents the possibility to break polymeric chains at very high shear.The main assumption of this kind of method is to assume the velocity profiles. We choose to solve the Modified Phan- Thien and Tanner model. The sum of the solvent viscosity and the polymeric viscosity constitutes the fluid viscosity . is the polymer relaxation time. The fifth term represents the variation of the viscosity in the contact. the polymeric stress tensor needs to be known to obtain the pressure field. 2. or segments. = 0 . a Modified Reynolds equation is obtained: ℎ + ℎ [5] ℎ ℎ = 12 +6 ℎ ℎ − 2ℎ + + + 2ℎ + + ℎ − ℎ − + +2 + − 12ℎ + ℎ ℎ + 2ℎ − + − +ℎ − + − with = ℎ and = ℎ the averaged Poiseuille velocities respectively in the radial and tangential directions. In the case of multigrade engine oils.2 Modified Phan-Thien and Tanner model The MPTT model considers the fluid as a Newtonian solvent including a homogenous and continuous network that has a non-Newtonian behavior (2). = . The polymeric chains. The third and fourth terms account for the inertia effects and the surface acceleration. The more numerous the loose ends. = 0 at = 0 and = 0 . and ξ are dimensionless parameters respectively 451 . we superpose a Couette flow and a Poiseuille flow determined without inertia terms. (−) being the trace function. the solvent corresponds to the basic mineral oil whereas the polymeric network stands for the long polymeric chains of the additives improving the VI. Proceeding to similar arrangements to Tichy and Bou-Saïd developments on the averaged momentum and continuity equations. The difference with the classical Reynolds equation stands in the second part of the right side member. interact between each other at their junctions. Finally. These junctions are not permanent but are continuously being created and destroyed and thus can result in loose end segments. At this state of the development. The global behavior of such a fluid is non-Newtonian.8. = 0 at = ℎ). That is practically illustrated by shear thinning effects. − − . the smaller the viscosity. the polymeric contribution to the pressure field stands in the two last terms. − . In Eq. Given the boundary conditions ( = . The MPTT model can be expressed as (6): = + [6] = ( + )= [7] 1 1 + . It avoids Calculation of Modified calculation instabilities at the outlet of the Reynolds equation thrust bearing.3 Energy equation The equation for the thermohydrodynamic (THD) model is a 3D thin film energy equation (8) that describes the heat transfer in the film thickness: + + + = + + [11] In order to avoid instabilities during the Initial conditions (t=0) numerical calculation. and therefore parameters fitting. To deal with the shear-thinning effect. on the final mixture of the lubricant. the temperature is calculated everywhere in the contact using the energy equation. three parameters. The pressure and Performances New rotational calculation the temperatures in the film thickness are speed and/or film thickness determined by using the over-relaxed t = t + dt (t<tend) Gauss-Seidel method. . . Obtain pressure field 2. we apply these two viscosity laws on the global viscosity of the oil. we introduce Cross equation (3): − = + [9] 1+( / ) where is the dynamic viscosity. Although only the polymeric part of the lubricant should suffer shear-thinning while the temperature should mainly affect the solvent part. Several models such as the Upper Convected Maxwell model and the White-Metzner model can be found when setting and ξ to specific values (7). it was only possible to perform rheological measurements.standing for the elongation and the slip between the polymeric network and the solvent. A flow chart of the global Obtain temperature field method. A Convergence Convergence? NO substitution method is used to solve the on pressure? MPTT model and provide the non- Newtonian stress tensor. To apply the Walther and McCoull model. the Walther and McCoull model was chosen (4): = 10 − [10] where is the kinematic viscosity under no shear rate. and the viscosities respectively under no and infinite shear rate. the critical shear rate and a parameter. Thus. As for the viscosity dependency on the temperature. It consists of calculating the temperature by using Assume velocity profiles forward or backward finite differences depending on the flow direction. 2. the temperature. the Richtmyer method (9) is introduced. Indeed. Solve MPTT model eight cases are implemented according to the velocities signs combinations. consists in Viscosity update an iterative process between velocities and pressure field calculations. near by the outer radius. Figure 1: Flowchart of We can then calculate the global bearing numerical solution 452 . presented in Figure 1.4 Numerical resolution Calculation of energy equation All equations are discretized using finite differences. The convergence criteria is reached when the maximum of the local pressure ratio is lower than 10-4. 1.02 0. we have = 7.03s-1 to 3000s-1. we use the last set of parameters.80. = 7. These rheological experiments demonstrate that Walther and McCoull model is a very accurate and robust model to calculate the viscosity as a function of the temperature.1. which is equivalent to the steady state case. An Anton Paar Physica MCR 301 rheometer with co- axial cylinders creating a Couette flow has been used for the test.03.04 0. The resolution of this problem is achieved using Fortran 90 and presented for the first time step.007 0.03 0. In-house experimental data Castrol exprimental data Walther & McCoull model based on three in-house experimental data Zoom-in between 20°C and 80°C Zoom-in between 100°C and 150°C 0. Unfortunately.1 Viscosity-temperature The viscosity of the lubricant was measured for temperatures going from -20°C to 150°C.91 and = 3.9).s] Viscosity [Pa. Current turbochargers speeds can reach up to 300000 rpm which correspond to shear-rates of 107s-1.025 0. In the numerical programs presented further in this paper.008 0. It is of course even better when its coefficients are based on three experimental data points. particularly for high temperatures. By fitting three experimental points using the least squares method. Based on the two data points from the oil manufacturer and by assuming that = 0.05 0. we obtain = 0.015 0.s] 0.performances and proceed to the next time step calculation.035 Viscosity [Pa.005 0.1 Rheological measurements Three types of rheological measurements were performed on the Castrol SLX Longtec Longlife II engine oil. the fluid was submitted to a shear rate ramp from 0. Averaged and repeated measurements were compared to Walther and McCoull model based on the two data points given by Castrol and based on three experimental data points (see Fig.005 0.004 0. For each intermediate temperature.2 Shear-thinning effect Some measurements were performed to obtain the parameters constituting the Cross equation (Eq. 3 EXPERIMENTAL STUDY 3.01 0.045 0.80 and = 3. the highest shear-rates 453 .07 which remain near by the first triplet.003 40 50 60 70 80 90 100 100 110 120 130 140 150 Temperature [°C] Temperature [°C] Figure 2: Castrol SLX Longtec Longlife II (0W30) Viscosity – temperature model correlation (a) Zoom-in between 40°C and 100°C - (b) Zoom-in between 100°C and 150°C 3.2).58. 3.006 0. 3. we only could observe the beginning of the shear-thinning effect. For that reason. which is a little bit far from the rotating frequency of turbochargers. respectively are [500 – 3000 Hz] and [500 – 2500 Hz]. The thrust bearing to be tested is seating in a cartridge.3) developed in house is used to compare experimental data to the prediction model.provided by the rheometer we used could not exceed 3000 s-1. we got data in that range.1. the oil inlet temperature or the oil inlet pressure are being under control. start-up events taken aside. Although those results tend to confirm that the relaxation time of engine oils is between 10-3 and 10-6 s (13-15). Results are summarized in Figure 5. the thrust oil film thickness is measured by using proximity probes.2 Thrust bearing rig A thrust bearing rig (see Fig. Using different oil than Castrol SLX Longtec Longlife II in simulations would be critical and need specific tests to be conducted too. But the oil having a very low viscosity and the machine being limited to 100Hz. 3. the axial load is provided by a pneumatic actuator and measured by a force transducer. Speed ranges of the studied thrust bearings A and B. 454 . Thus the parameter ∞ standing for the viscosity at infinitely high shear-rate and which is the critical shear-rate standing for the transition between the high and the low viscosity plates have been obtained from literature (10) for this oil submitted to turbocharger solicitations in term of shear rate magnitude. a second load cell is implemented to measure the thrust oil resistive torque (and consequently thrust power loss) and many other operating parameters such as the rotational speed.4) operating in steady state. Therefore. the simulation conditions presented in Table 1 are relevant to turbocharger operating conditions. Actuator Load cell Cartridge Driver Figure 3: Thrust bearing rig 4 PARAMETRIC STUDY – EXPERIMENTAL COMPARISON Correlation between the prediction code and test data in term of axial load and power loss are presented for two typical turbocharger thrust bearings (see Fig.3 Relaxation time To access the relaxation time of the lubricant. we used the oscillatory mode of the Anton Paar rheometer to run a frequency sweep and obtain its loss and storage moduli (11-12). The energy equation is solved to provide the temperature at each node of the mesh. With the assumptions made. Figure 4: Thrust bearings tested on the rig for numerical correlation Table 1: Tested thrust bearings geometries and operating conditions Supply Supply Number ID OD Speed Oil pressure temperature of pads [mm] [mm] [Hz] Grade [bar] [°C] Bearing 6 8.2 Case 2: 100% solvent – THD analysis In this case.1 Case 1: 100% solvent – Effective temperature For this case. This may come from the fact that conduction is not calculated in the solids and adiabatic boundary conditions are used for the fluid-solid interfaces. not using the MPTT model).5). the MPTT model is not used (i.6 14. Indeed. This phenomenon occurs on the thrust bearing rig.65 1700 0 100 B 10W30 The prediction model takes into account operating condition of the two sides of thrust bearing.15 2500 0 120 0W30 A Bearing SAE 4 13. 455 .25 19. The predicted axial loads correlate with the experimental data obtained on the thrust bearing rig. The consideration of the conduction problem in the solids could possibly reduce the discrepancy observed at high axial loading. In this thermohydrodynamic analysis. while the active side of the bearing works with a certain thickness. The discrepancy between experimental data and predictions in terms of axial load narrows but still exist nearby 15% at low film thickness (see Fig.5). the predicted axial load remains acceptable but we observe a discrepancy for power loss that needs to be investigated. 4. the code assumes that the oil pockets of the bearing provide fresh oil at 120°C. 100% Newtonian solvent is considered) but temperature effects on the viscosity are still taken into account.e. enabling the update of the viscosity. the code is run considering that the lubricant is only made of solvent (i. the remaining clearance leaves a determined thickness on the other side of the bearing which creates an axial load in the opposite direction and additional power loss. 4. especially when the film thickness increases (see Fig. as well as on engine.e. 6 0. This gap with the experimental data points 456 .2 0.7 0. predictions Because of the very small size of turbocharger thrust bearings. Therefore we decide to benchmark our code using an experiment done for large bearing by Dadouche (16) and Ahmed (17). 10-6].4 0.6 0.05 and we consider that no polymer chain breaks ( = 0 ). the power loss is overestimated when the thickness is lower than 15 µm.5 0.8 0.98% solvent.6 1 0. 2% polym ers Power loss . (13) who considered that the lubricants followed a Maxwell model as the relaxation times they found were in the interval [10-5 .Averaged experim ental thrust bearing rig data Axial load prediction -THD .8 0.3 0.7 1 0. Bearing A Bearing B 1.6). we compare their experimental data with our predictions.4 1. we set the relaxation time to 5. However.10-6 s.4 0.2 0.9 1. For the MPTT model. oil properties and operating conditions.1 0.4 0. It suits the analyses performed on several multigrade oils by Bates et al.2 0.8 ~15% 0.5). This is a realistic assumption when discussing with oil manufacturers (18).9 0.4 1 1.98% solvent.8 1. The slip parameter ξ is set to 0.6 0.100% solvent Power loss prediction -THD .3 0. Figure 6: Temperature correlations with Dadouche (16) and Ahmed (17) experiments 4. Based on their geometry. 2% polym ers Figure 5: Axial load and power loss correlations: Test data vs. the code predicts perfectly the axial load (see Fig.2 0.1 0 0 0 0 7 12 17 22 27 7 12 17 22 27 Oil film thickness [µm] Oil flim thickness [µm] Axial load .2 0.6 0. They are within the range of 4% which is really acceptable and validates the temperature model (see Fig.2 Total axial load (dimensionless) Total axial load (dimensionless) 1 Power loss (dimensionless) Power loss (dimensionless) 0.3 Case 3: 98% solvent – THD analysis We now consider that the engine oil is made of 98% of solvent and 2% of polymers. In this case.4 0.100% solvent Axial load prediction -THD .5 0. it is impossible to put thermocouple in the pad without disturbing the pressure field.Averaged experim ental thrust bearing rig data Power loss prediction -THD . Investigations on high speed torque meters are being considered as well. 1977 (3) CROSS M. et al. J. New York. It is important to understand whether a quasi steady state approach is sufficient to model a transient event or the unsteady solution is required. Numerous parameters for the viscosity models are based on rheological experiments performed on Castrol SLX Longtec Longlife II (0W30) but device limitations prevent us from obtaining a precise relaxation time and the remaining properties were set from the literature or industrial background. – Network Theories for Polymer melts and Concentrated Solutions. For those two bearings. Texas A&M University. p. Paris. Dynamics of Polymeric Liquids.. 1987 (7) EHERT P. Collection de la Direction des Etudes et Recherches d’ED. Those discrepancies may come from inaccuracies in the mass flow rate determination and the adiabatic hypothesis for the film temperature at the fluid-solid interfaces. An additional in-house system based on the enthalpy measurement (19) is to be mounted on the rig to double check this quantity. Thèse Mécanique. 1965 (4) ASTM D341-09. power loss. Mech. Elsevier. Journal of Non-Newtonian Fluid Mechanics.M. The target is to getting right the first time with product design prediction tool. 1993 (8) FRENE J. On the other hand. INSA de Lyon. REFERENCE LIST (1) WILLIAMSON B. 1967 (10) SAN ANDREAS L. – Difference Methods for Initial-Value Problems. ASTM International. – Internal Testing Report. – The Viscoelastic Properties of Multigrade Oils and their Effects on Journal-Bearing Characteristics. 73. New York. Paliers et butées. p. BOU-SAÏD B. – A New Constitutive Equation Derived from Network Theory.P. 488 p.505-512.D. Vol. p. Non-Newtonian Fluid Mech. Second Edition. et al. John Wiley & Sons. Rheological tests in high shear and high frequency ranges would be a further investigation on the parameters’ accuracy for this oil. J. 20(5). Interscience Publications.B. Some investigations. – Contribution à l’Etude du Comportement de Mécanismes Lubrifiés sous Chargements Transitoires. 5 CONCLUSION A numerical model is being developed and uses a thermohydrodynamic analysis coupled with the MPTT model and two viscosity laws. the precision of the power loss measurement device can be questioned because of the very low friction torque to measure. such as fluid recirculation or under-lubrication. As for the power loss estimations. Eyrolles. – Rheology of Non-Newtonian Fluids: A New Flow Equation for Pseudoplastic Systems.417-437.. 2009 457 .W.narrows as the thickness increases and the predicted power loss finally tends to stabilize around better values than previous predictions. Tribology Transactions. 1991 (6) BIRD R. using adequate parameters for the MPTT model demonstrates a very good correlation between the predicted axial load and experimental data. – Hydrodynamic Lubrication and Bearing Behavior with Impulsive Loads..A.34(4). 1997 (2) PHAN-THIEN N. the prediction doesn’t match very well with the measured values. Eng. 2009 (5) TICHY J. 1990 (9) RICHTMYER R. – Lubrification hydrodynamique.. Dept.115-126.. 246p. TANNER R. oil flow…) at the development stage of turbocharger. It has been presented for the first time step but is under development for the unsteady formulation.I. et al. are being considered to improve the predicted torque and therefore the power loss. MORTON K. The developed model will be used for the optimization of thrust bearing design performances (thrust load capacity. Colloid Sci. Standard Practice for Viscosity-Temperature Charts for Liquid Petroleum Products. p. June 8-12 2009. Orlando. – Determination of Discrete relaxation and Retardation Time Spectra from Dynamic Mechanical Data. Sea and Air. 2008 (18) AYEL J..511-519. – Etude des Phénomènes Thermiques dans les Butées Hydrodynamiques. WALTERS K. ASTM International.. Technique de l’Ingénieur. GJIKA K. Endeavour. p. Rheologica Acta. USA. 195p. 28. BM5354. FL. et al.. 15p. 1986 (14) CROCHET M.A. – A Correlation Between Engine Oil Rheology and Oil Film Thickness in Engine Journal Bearing.H. Thèse Mécanique. – Calculation of Relaxation Spectra from Mechanical Spectra in Matlab.64-77. 136p.(11) KONTOGIORGOS V.. 2010 (12) BAUMGAERTEL M. WINTER H. SAE Technical Paper Series.. – Computational Rheology : A New Science. 2002 (19) LAMQUIN T. GT2009-59599 458 . et al. p. Polymer Testing 29. – Contribution à l’Etude des Effets Thermiques et des Déformations dans les Butées Hydrodynamiques à Géométrie Fixe. 1989 (13) BATES T. – Power Losses Identification on Turbocharger Hydrodynamic Bearing Systems: Test and Prediction. Proceedings of ASME Turbo Expo 2009: Power for Land. Thèse Mécanique Productique Transport.1021-1025. 1977 (16) DADOUCHE A. 1998 (17) AHMED S. 108 p. – The Relationship between Engine Oil Viscosity and Engine Performance. 17:2.. – Lubrifiants – Additifs à action physique ou physiologique. 1993 (15) LONGSTRUP T. Université de Poitiers. could lead to thermomechanical fatigue (TMF) of the manifold and the turbine housing. UK There is a high demand from automotive regulators to reduce emissions of air pollutants such as oxides of nitrogen (NOx) and oxides of sulphur (SOX). Experimental methods are used to measure temperature in the housing and to validate the FEA results. whereas by contrast a truck travelling on the highway will have low frequency of cycles. There are other factors which could reduce the TMF life of the vehicle. Different vehicles have different duty cycles. as it can be observed in this application the frequency of the cycles where the temperature changes from minimum to maximum is high when compared with the cycles shown in Figure 2 measured from vehicle ‘B’. 2014 461 . for example increase in altitude above the sea level increases the operating temperature of the turbine housing and hence a reduction in TMF life. for example forklift trucks have very cyclic applications. This increase in temperature combined with thermal cycles with different amplitudes to which the vehicle is subjected. Finite element analysis (FEA) is used to analyse the mathematical model of the turbine housing and evaluate the temperature and strain distribution in the housing. This paper examines a variety of thermal duty cycles obtained from different vehicles and considers them for analysis. One solution is to increase the exhaust gas temperature to start chemical reactions in the exhaust manifold. NOTATION K= Thermal conductivity [W/mK] Cp= Specific heat capacity [J/kgK] ρ= Density [kg/m3] FO =Fourier Modulus [Dimensionless] t= Time [s] x=Length of the element [m] 1. Two such cycles are given in Figure 1 and Figure 2. INTRODUCTION During its life time in service the turbine housing of a turbocharger is subjected to different thermal cycles.Influence of holding time in thermal cycle proven by state-of-the-art thermomechanical fatigue calculation in the turbine housing of a turbocharger K Shoghi BorgWarner Turbo Systems Ltd. One cycle based on the results of the analysis is selected for detailed study. This is as expected since in vehicle ‘A’ application the torque output from the engine can vary significantly depending on the number of operations where a load is lifted or lowered whereas in vehicle ‘B’ the torque is _______________________________________ © The author(s) and/or their employer(s). The thermal cycles shown in Figure 1 are measured on vehicle ‘A’. However. However in the gas stand test. The thermal cycle test performed on the gas stand offers more control regarding temperature and the turbine stage can operate independently of the compressor stage due to the fact that in a gas stand test the output of the compressor stage does not influence the inlet of the turbine stage as it does in an engine test. vehicle ‘A’ has a significantly lower mean temperature than vehicle ‘B’. Figure 2 Thermal cycle measured for the application with vehicle ‘B’. the effect of pulsation is not simulated as it would be the case on the engine test bed. Thermal cycle testing on a turbine housing can be performed either on the engine test bed or gas stand test rig. under gas stand test condition these ramp-ups and cooling down in temperatures 462 . Figure 1 Thermal cycle measured for the application with vehicle ‘A’.approximately constant. The investigation carried out here considers a combination of the thermal cycles ‘A’ and ‘B’. Although there may be rapid spikes in temperature as shown in Figure 1. It is assumed that Fourier modulus. and step time used for the analysis which can influence the values of plastic strain in the model. The heat transfer coefficients were calculated using both 1-D and 3-D approach. consisting of cylinder head. Δx in direction of heat flow can be calculated using equation(1). As it will be demonstrated in experimental work. Figure 3 Typical test cycle used at the gas stand. The minimum time step. nominal load. element size considered. 2. A typical test cycle is presented in Figure3 showing acceleration between 1-2. Y. Hence all degrees of freedom in X.are not very practical to implement. F0 =1 as the components are almost in thermal equilibrium with their surroundings [2]. The coupled transient heat transfer and structural model was constrained as shown in Figure 4. exhaust manifold. and Z directions can be constrained using the surface and the constructed lines on the 463 . Some engine manufacturers may have their own test cycle with different acceleration. it is important to allow the growth of the structure from the central position. For the 3-D method. CFD code was used to calculate the heat transfer coefficients and the solution was coupled directly to structural FEA code. deceleration. FINITE ELEMENT AND COMPUTATIONAL FLUID DYNAMICS MODELS The finite element methodology for a simplified model for thermomechanical fatigue was discussed by Oberste-Brandenburg et al [1]. Δt for the largest element size. In order to prevent over constraining the turbine housing and the manifold. and idle time allocated. nominal load from 2-3 and idle from 4-5. 4KΔt F0 = 2 (1) ρcp (Δ x) The model is constructed and meshed as shown in Figure 4. a second cycle was selected to allow for longer time for heat soak. The methodology of 1-D approach was discussed in [1]. This test cycle offers the higher temperatures as shown in the thermal cycle ‘B’ and a moderate ramp-up as thermal cycle ‘A’. as these will result in artificial stress concentration at these locations. However FEA can simulate these conditions where the time between maximum and minimum temperature is between 10-15 seconds. However one important aspect is the relationship between the ramp-up and ramp-down of the cycle. turbine housing and bearing housing. The test was performed at two different gas stands due to different test constraints such gas flow rate and temperature control. The mesh of each fluid and solid part consisting of: rotating fluid (the space between the turbine blades).cylinder head to prevent rigid body motion without over constraining the model. For example transient CFD coupled with steady state heat transfer and structural analysis or conjugate heat transfer coupled directly with structural analysis and finally steady state CFD analysis could be used with different time frames combined with transient heat transfer and structural analysis. with temperature dependent material. Figure 5 Giving temperatures in the turbine housing for the time steps at 300 and 800 seconds. There are different ways in which the fluid model can be coupled with the structural model to obtain the same results. was used. The total thermal energy equation with a turbulent k-ε model was used for the analysis. bearing housing. multilinear kinematic hardening model. The boundary condition of the model is defined by fixing the mass flow rate at the inlet to the turbine housing and the average static pressure at the outlet which determines the pressure ratio required for the solution. manifold. To account for the accumulated plastic strain in the turbine housing at different temperatures. stationary fluid (the gas passage of the turbine housing and manifold) and solid (turbine housing. This is the option which has been considered due to the computing cost and speed compared with other methods. 464 . and cylinder head) was imported into the CFX separately. The Ansys CXF code was used to construct the CFD model of the turbine housing assembly. Figure 4 Showing the mesh and constraint of the turbine housing assembly. The cycle shown in Figure 3 was used in the transient heat transfer model. It is worth noting that these cycles are designed for use with different gas stands which may have different characteristics. 465 . As shown in Figure 6 the FEA results have taken more time for the turbine housing to cool down. the temperature is still climbing after 400 seconds and as the strength of the turbine material decreases with the rise in temperature. Figure 6 Showing the thermal cycle output from an FEA file. This is given in Figure 7. The corresponding plastic strain was obtained from the structural part of the model as the criterion for crack initiation and failure is based on the limit of this plastic strain.The corresponding temperature for different time steps in the cycle from the transient heat transfer analysis is given in Figure 5. a second cycle was considered with a ramp-up time of 200 seconds as this cycle offers more time for the heat soak when compared with the 100 second cycle. Figure 7 Showing the thermal cycle with a ramp-up of 200 seconds. In order to simulate the effect of heat soak. but not be possible with another. This is due to the thermal inertia stored in the turbine housing assembly. the 100 seconds ramp-up cycle would not be suitable for this test. unless the heat soak time was increased. As it shall be demonstrated in experimental work. A cycle with a ramp-up time of 100 seconds may be possible on one gas stand. Mass flow Temperature control control Figure 9 Actual test cycle on a gas stand. The test was performed on a gas stand using the cycle shown in Figure 3. It was necessary to increase the time of the cycle to achieve the temperature required at the end of the cycle. 466 . In order to see the effect of heat soak a cycle with the ramp-up of 200 seconds was tested on a gas stand as discussed in section 3. The results are shown in Figure 10. then it is possible to simulate the desired cycle. initially during the ramp-up the cycle was controlled by means of mass flow through the inlet up to a nominal value. EXPERIMENTAL WORK AND GAS STAND TESTING The tests were performed on two different gas stands. the final temperature of the cycle was controlled by means of temperature control as given by Figure 9.3. the true values of the strain in the structure will not be representative of the true nature of the cycle. One with capability of using a shorter ramp-up time while the other requires longer ramp-up and ramp-down times. The cycle given in Figure 3 was simulated during the gas stand test. After 1200 cycles the housing was examined for cracks. It is worth stating that if a tolerance on the final temperature is allowed. It is important to note that the average temperature of the turbine inlet gas is higher than the metal temperatures measured by the thermocouples. however in order to prevent the final temperature of the cycle from exceeding its intended maximum value. The turbine housing was fitted with thermocouples as shown in Figure 8. Figure 8 Positions of thermocouples on the turbine housing. Also that the maximum gas temperature for each thermocouple occurs at the same time interval because if the cooling stage of the cycle starts before the maximum temperature is achieved. As there are two thermocouples fitted on the inlet flange of the turbine housing (please refer to Figure 8 in section 3). This result is consistent with the temperature measurements on the previous cycle. It is due to the fact that the tongue heats up much quicker than the rest of the turbine housing and the resulting temperature difference combined with the geometry of the volute results in crack initiation at this location. VALIDATION The comparison between the FEA thermal cycle with the ramp-up of 200 seconds and the experimental work on the gas stand with the same ramp-up time for position 1 is given in Figure 11. 467 . It can be observed that there is a good correlation between the experimental work and the FEA model. The FEA model for the more severe cycle given in Figure 3 discussed in section 3 with the ramp-up of 100seconds is shown in Figure 12. this is the case in majority of the applications. 800 700 600 TC1 Temperature °C 500 TC2 400 TC4 TC6 300 TC7 200 TC8 100 Avrg at turbine inlet 0 0 500 1000 1500 Time (Seconds) Figure 10 Test cycle obtained from a gas stand test. The correlation between the high strain locations obtained from the FEA model and the turbine housing section is given in Figure 13.Bist et al [4].Laengler et al [3] and S. this position was selected for the comparison with the FEA results. It is interesting to note that the values for the temperature from the FEA model measured at the inlet flange of the turbine housing are nearly the same. The housing was sectioned along its circumference through the divider wall between the scrolls in the volute. It is important to observe that the mode of failure starts at the tongue in the turbine housing. However it is important that the crack does not propagate to the outer skin as this will result in leakage of exhaust gas through the housing and loss of performance and failure. The study of fatigue crack growth and failure prediction is discussed by F. 4. Figure 12 FEA for the thermal cycle with 200 seconds ramp-up at 600 seconds. Figure 11 Comparisons between test and FEA for the thermal cycle with 200 seconds ramp-up. Figure 13 Comparisons between test and FEA. 468 . Gugau.Kannusamy.There was good correlation between experimental results and theoretical work. S. 19-20 May 2010. C.Mao.86-93.A. Transient Thermal Analysis Nov 2009.Two different thermal cycles were simulated. F. On the influence of thermal boundary conditions on Thermo Mechanical Analysis of turbine housing of a turbocharger. chapter 3. London. P47. pp. Validation of a phenomenological lifetime estimation approach for application on turbine housing of turbochargers.Laengler.193-205.SUMMARY 1. pp.207-215.Both F. P. 15-16 May 2012. 9th International Conference on Turbochargers and Turbocharging. H. 2. 19-20 May 2010. 2. 3. London.Liang. Thermomechanical fatigue crack growth and failure prediction for turbine housings. pp. Further test on engine test bed is recommended and investigation into effect of pulsation in the turbine housing assembly. and F. Experimental work is planned to measure the values of strain in the housing by using: (i) Strain gauges (ii) Neutron diffraction methodology (this method is ideal for the locations where it is not practical to fit a strain gauge).As it was demonstrated by experimental methods. 2. 9th International Conference on Turbochargers and Turbocharging. A.Alesksanoglu. 469 . E.Kruse. 3. T. 4.Shoghi. CONCLUSIONS AND FUTURE WORK 1.The FEA model successfully predicted the mode of failure and this matched very well to the location of cracks developed in the turbine housing during the durability test.Oberste-Brandenburg. K. M. London. REFERENCES 1. Ansys theory manual.Scholz. FEA and CFD codes there must be enough time allowed for the heat soak during the thermal cycle test.Bist. R. and CFD methodology were used to simulate thermal cycles in a turbine housing assembly of a turbocharger.Tayal.E. 10th International Conference on Turbochargers and Turbocharging. otherwise the full effect of the damage due to thermal stresses and strains in the structure will not be captured. 11th International Conference on Turbochargers and Turbocharging . Combustion Engines & Fuels Group Organising Committee Dr Kian Banisoleiman (Chairman) Lloyd’s Register Dr Roland Baar Technische Universität Berlin Andrew Banks Ricardo Steve Birnie BorgWarner Dr Chris Brace University of Bath Dr Geoff Capon Ford Dr Ennio Codan ABB Gavin Donkin Honeywell Dr-Ing Dietmar Filsinger IHI Pierre French Cummins Turbo Technologies Dr Seiichi Ibaraki Mitsubishi Per-Inge Larson Scania Prof Ricardo Martinez-Botas Imperial College London Takashi Otobe Honda Prof Joerg Seume Hanover University Dr Les Smith Jaguar Land Rover Prof Stephen Spence Queens University Belfast Bernd Wietholt Volkswagen . LONDON Conference Proceedings sponsored by: AMSTERDAM BOSTON CAMBRIDGE HEIDELBERG LONDON NEW YORK OXFORD PARIS SAN DIEGO SAN FRANCISCO SINGAPORE SYDNEY TOKYO Woodhead Publishing is an imprint of Elsevier .11th International Conference on Turbochargers and Turbocharging 13–14 MAY 2014 THE BRITISH MUSEUM. Waltham. 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